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Vehicle Design 


Advances in Vehicle Design 

To Ruth 

Advances in Vehicle Design 


John Fenton 




Professional Engineering Publishing Limited 
London and Bury St Edmunds, UK 

First published 1999 

This publication is copyright under the Berne Convention and the International 
Copyright Convention. All rights reserved. Apart from any fair dealing for the 
purpose of private study, research, criticism, or review, as permitted under the 
Copyright Designs and Patents Act 1988, no part may be reproduced, stored in 
a retrieval system, or transmitted in any form or by any means, electronic, 
electrical, chemical, mechanical, photocopying, recording or otherwise, without 
the prior permissionof the copyright owners. Unlicensed multiple copying of this 
publication is illegal. Inquiries should be addressed to: The Publishing Editor, 
Mechanical Engineering Publications Limited, Northgate Avenue, Bury St 
Edmunds, Suffolk, IP32 6BW, UK. 

© John Fenton 

ISBN 1 86058 181 1 

A CIP catalogue record for this book is available from the British Library. 

The publishers are not responsible for any statement made in this publication. 
Data, discussion, and conclusions developed by the Author are for information 
only and are not intended for use without independent substantiating investi- 
gation on the part of the potential users. Opinions expressed are those of the 
Author and are not necessarily those of the Institution of Mechanical 
Engineers or its publishers. 

Printed and bound in Great Britain by St Edmundsbury Press Limted, 

Suffolk, UK 


vii Preface 

Chapter 1 : Materials and construction advances 

2 Steel durability and structural efficiency 

3 Vigorous development of light alloys 

4 Hybrid metal/plastic systems 

7 Recycled PET, and prime PBT, for sun-roof parts 
10 Material property charts and performance indices 
13 Design for self-pierce riveting 

Chapter 2: Structure and safety 

20 Structure analysis for interior noise 
24 Preparing for statutory pedestrian protection 
27 Design for the disabled 
3 1 Adaptive restraint technologies 

Chapter 3: Powertrain/chassis systems 

36 Powertrains: the next stage? 

45 Constant-pressure cycle: the future for diesels? 

47 Valve arrangements for enhanced engine efficiency 
52 Trends in transmission design 
58 The mechanics of roll-over 
62 Suspension and steering linkage analysis 

Chapter 4: Electrical and electronic systems 

70 Automotive electronics maturity 
76 Navigation system advances 
79 Digital circuits for computation 
8 1 Proprietary control system advances 
84 Hybrid drive prospects 
90 Automation of handling tests 

Chapter 5: Vehicle development 

96 Mercedes A-class 
102 Ford Focus 
108 Land Rover Freelander 
1 12 Project Thrust SSC 

Chapter 6: Systems development: powertrain/chassis 

118 Engine developments 

122 Engine induction systems 

124 Refinement and reduced emissions 

128 Drive and steer systems 

133 Suspension development 

139 Braking systems 

Chapter 7: System developments: body structure/sy stems 

144 Structural systems 

144 Controlled collapse 

145 Body shell integrity 

147 Chassis/body shell elements 

15 1 Car body systems 

151 Occupant restraint 

155 Doors, windows and panels 

159 Trim and fittings 

161 Aerodynamics and weight saving 

163 CV systems 

163 CV chassis-cab configuration 

164 Cab/body fittings 

168 Advanced bus/ambulance design 

173 References 
175 Index 


This literature survey is aimed at providing the vehicle design engineer with an 
update in vehicle and body systems. The author has scanned the considerable 
output of technical presentations during 1997- 9 to extract and distil developing 
technologies of particular import to the working designer. The easily digestible 
presentation, with unusually high dependence on diagrammatic presentation, 
continues the popular style used for the original handbooks that were compiled 
by the author, and published by Professional Engineering Publishing. These are 
listed on the Related Titles page overleaf. Advances in Vehicle Design serves both 
as an update to the earlier volumes and as a stand-alone volume. The referenced 
leads provided in the text are intended to help designers and engineers from 
whatever backgrounddiscipline. Widespread availability of computing power to 
designers and engineers has created the possibility of considerably shortening the 
lead-times between design conception and prototype manufacture. Much of the 
material covered here will assist in establishing predictive techniques. 

Advances in Vehicle Design is an update of vehicle and body systems design 
in that it provides readers with an insight into analytical methods given in a wide 
variety of published sources such as; technical journals, conference papers, and 
proceedings of engineering institutions, for which a comprehensive list of refer- 
ences is provided. The analyses are therefore not necessarily fully developed or 
rigorously evaluated. Recourse to the original references is necessary particularly 
in order to understand the limiting assumptions on which the analyses are based. 

Much of the analytical work is centred around impending legislation and, where 
this is quoted in the text, it is for illustration only and it is, of course, important to 
examine the latest statutes in the locality concerned. The list of references given 
at the end of the volume is a key element of the publication, providing where 
possible a link to the original publication source. Where the original publication is 
not available in bookshops, many of the sources can be found in libraries such as 
those of the Institution of Mechanical Engineers, London, or the Motor Industry 
Research Association, in Nuneaton, UK, as well astheBritishLibrary.Othersimilar 
respositories of techincal information should be able to provide a selection of 
original source material. Where the source is a company announcement of 
techniques and systems, names, but not addresses, of the companies/consultan- 
cies are given. Most operate internationally and have different national locations, 
best found by enquiry in the country concerned. For the patent reviews in chapters 
six and seven, full specifications can be purchased from The Patent Office, Cardiff 
Road, Newport, NP9 1RH, UK. 


Related Titles 




Multi-Body Dynamics 


1 80058 122 6 

Gasoline Engine Analysis 

J Fenton 

0 85298 634 3 

Handbook of Automotive Body Construction 

J Fenton 

1 86058 073 4 

and Design Analysis 

Cranes - Design, Practice and Maintenance 

J Verschoof 

1 86058 130 7 

Handbook of Automotive Body Systems Design 

J Fenton 

1 86058 067 X 

Handbook of Automotive Powertrain and 

J Fenton 

1 86058 075 0 

Chassis Design 

Vehicle Handling Dynamics 

J Ellis 

0 85298 885 0 

Handbook of Vehicle Design Analysis 

J Fenton 

0 85298 963 6 

Automotive Braking: - Recent Developments 

D Barton 

1 86058 131 5 

Brakes and Friction Materials 

G A Harper 

1 86058 127 7 

Automotive Engineer Monthly Magazine 

W Kimberley 


Journal of Automotive Engineering (IMechE Proceedings Part D) 


For full details of all Professional Engineering Publications please contact: 

Sales Department 

Professional Engineering Publishing 

Nothgate Avenue 

Bury St Edmunds 

IP32 6BW 


Telephone: +44(0)1284 763 277 
Fax: +44(0)1284 718 693 


Chapter 1 : 

Materials and construction advances 

The considerable comeback made by the steel industry in 
restating its case for structural superiority over the light alloys, 
and the ever moving goalposts of developing aluminium and 
polymer composites, open this chapter. The re-emergence of 
hybrid metal/plastic structures is also discussed as well as the 
creation of reinforced plastics with recycled-polymer matrices. 
The move to material property charts which lead to the 
creation of performance indices is next examined and the 
chapter concludes with the efforts to make self-pierce riveting 
a viable alternative to conventional welded joints in body 



Steel durability 

and structural efficiency 

According to researchers at Lotus Engineering 1 , mar- 
ket expectations for durability of vehicle body panels 
is typically seven years or 100 000 miles, with an 
expectation of 10 years corrosion-free life. British 
Steel engineers 2 have shown the main influences on 
durability with the corrosion triangle of Fig 1. In 
product design the traditional approach has been to 
remove moisture traps, allow better penetration of 
paint and evolve design with greater resistance to 
stone chipping. There had also been gradual adoption 
of one-side zinc-coated steel panels, offering protec- 
tion on the inside and good paintability on the out- 
side. More recently there had been a move from hot 
dip galvanized to electrozinc panels, spurred by 
Japanese manufacturers; these offer a range of 
ductilities, increased weldability and the ability to 
alloy the coating to provide various coating thick- 
nesses and corrosion resistance levels. Now double- 
sided coated steels are favoured with differential 
coating weights: typical thicknesses are now 45-60 
g/m 2 . Future quality improvements are promised by 
development work in surface treatment and 
formability. Permanent organic based topcoats or 
phosphate coatings are providing improved perform- 
ance in both weldability and corrosion prevention. 
Another new initiative is a drive by the OEMs to 
replace electroplated products by ‘galvanneal’ ones 
for outside parts, in the interests of production 
economy. These have the ability to provide good 

In a joint venture between British Steel Strip 
Products and Rover Group, trial parts were prepared 
for the Rover 600, including front fender, front door 
skin, rear door skin and bonnet. In 1997 all new 
models were produced with two-sided galvanneal 
full-finished panels. In the longer term, the authors 
see the development of non-bake hardening, higher 
strength steel substrates with enhanced formability 
for use with galvanneal coatings, and the introduc- 
tion of extra high strength transformation induced 
plasticity (TRIP) steels as substrates for zinc type 
coating which will offer yield strength approaching 
1500 MPa, alongside high ductility. Prepainted body 
finished sheet steel, which obviate the electropheritic 
primer coat are also expected, as is a possible new 
coating process based on vapour deposition of zinc in 

a vacuum. 

Car body weight reduction of 25%, without 
cost penalty, plus a 20% reduction in part count, has 
been the result of the final phase of the USLAB 
project for light-weighting steel automotive struc- 
tures carried out by an international consortium of 
steel companies supervised by Porsche Engineering 
Services. The objective of a feasible design, using 
commercially available materials and manufactur- 
ing processes, has also been met. An 80% gain in 
torsional rigidity, and 52% in bending, has also been 
recorded and estimated body-shell cost of $947 com- 
pares with $980 for a year 2000 comparison figure for 

Materials of 


(including micro environment) 



Fig 1: Triangle of factors affecting corrosion 

Fig 2: ULSAB body shell 



a conventionally constructed body to a similar speci- 
fication. The 203 kg body shell (Fig 2), of 2.7 metres 
wheelbase, has a torsional rigidity of 20,800 Nm/deg 
and a first torsional vibration mode of 60 Hz, using 
the well-documented construction techniques. 

A supercomputer analysis of crashworthiness 
has shown that frontal NCAP and rear moving- 
barrier levels have been achieved at 17% higher than 
the required speed. The body has also satisfied 55 
kph/50% offset AMS, European side impact and roof 
crush requirements. The 35 mph frontal NCAP showed 
a peak deceleration of 3 1 g, considered satisfactory in 
that stiffer body sides are required to meet 50% AMS 
offset requirements. The offset AMS frontal impact 
deceleration showed a peak at 35g, considered a good 
result in relation to the severity of the test. The 
simulation was carried out at 1350 kg kerb weight, 
plus 1 1 3 kg luggage and 149 kg for two occupants. At 
the final count, high strength steel was used for 90% 
of the structure, ranging from 210 to 550 MPa yield, 
in gauges from 0.65 to 2 mm. Around 45% of the 
structure involved laser-welded tailored blanks, in- 
cluding the monoside panels, Fig 3, which ranged in 
gauge from 0.7 to 1.7 mm and was made up of steel 
elements having yield strengths from 210 to 350 
MPa. Fig 4 shows the hydroformed side roof rail. 

To the casual observer the elements of the 
complete body shell ‘look’ more structurally effi- 
cient but they still seem a far cry from a fully 
optimized shape for the steel ‘backbone’, as much of 
the material seems to be still used in relatively low 
stress areas where its function is as a ‘cover’ rather 
than a fully working structural element. Perhaps the 
next stage in weight reduction should be in designing 
a highly efficient steel monocoque, then moulding 
over plastic elements for the non-structural covering 
and closure shaping details, using the hybrid metal/ 
plastics techniques discussed later in this chapter. 

Fig 5 Al , concept 
car and structure 


development of light alloys 

Phase 2 of Audi’s aluminium alloy body construction 
programme was unveiled at a recent Motor Show in 
the form of the Al 2 concept car, Fig 5. The 3.76 metre 
long times 1 .56 metre high car weighs just 750 kg in 
1.2 litre engined form, some 250 kg less than a 
conventional steel body vehicle, it was argued. The 
number of cast nodes has been reduced compared 
with the Phase-one aluminium alloy structure of the 
Audi A8. Most of the nodes are now produced by 
butt-welding the extruded sections. High level seat- 
ing is provided over a sandwich-construction floor. 
Use of internal high pressure reshaping techniques 
has reduced the number of shaping and cutting opera- 
tions required 

The 1997 International Magnesium Associa- 
tion Design Award for applications went to the die- 
cast instrument panel support beam. Fig 6, on the GM 
G-van. It is a one-piece design weighing, at 1 2.2 kg, 
5.9 kg less than the welded steel tubular structure it 
replaced. Proponents of magnesium alloy are point- 
ing out that as die-castings in the material solidify 
from the outside in, they enjoy a dense chilled skin 
together with a relatively course-grained interior. 
Skin thickness is said to be relatively constant regard- 
less of total wall thickness so that reductions in 
section size can often be made when the material is 
used as a substitute. The AZ91D (9% aluminium, 1% 
zinc) alloy is now finding application in removable 
rear seats for minivans. In the seat systems shown in 
Fig 7, Delphi Automotive have achieved 60% weight 
reduction over conventional construction by adopt- 
ing a cast magnesium cushion frame and an extruded 



aluminium back frame. Other magnesium grades are 
being used in combination with engineering plastics 
for composite instrument panels. VW have also 
demonstrated a Polo door in pressure die-cast magne- 
sium with a carbon fibre reinforced outer panel, 
which offers over 40% weight reduction over con- 
ventional steel construction. 

Hybrid metal/plastic systems 

Metal/plastic composites have been promoted again 
by Bayer in a recent presentation to body engineers 3 . 
A compelling argument is made for using simplified 
thin-wall structures in high strength metals which are 
stabilized by plastic composites which can also cut 
down on welding operations when combined as an in- 
mould assembly. The complexity of CV-cab and car 
bodies has often ruled against the effective use of 
lightweight metal systems on their own. 

In the latest version of the technique, the pro- 
cesses of inserting, where metal parts such as bushes 
are included in a polymer moulding, and/or outselling, 
where various functions in plastic are moulded on to 
a metal baseplate, are taken a stage further. Cross- 
sectional distortion of thin-walled metal beams can 
be prevented by relatively small forces applied in the 
new process by the presence of moulded plastic 
supports in the form of x-pattem ribbing. 

The deformation behaviour of various section 
profiles under bending and torsion is shown in Fig 8 
a and b. Interconnecting points between plastic and 
metal are preformed in the metal part before it enters 
the plastics mould. Either corrosion-protected steel 
or aluminium alloy is the normal choice of metal with 
glass-fibre reinforced, impact modified, polyamide- 
6 (Bayer’s Durethan BKV) being the plastic choice. 
Bayer say it is not always desirable to have the metal 
‘preform’ in one piece, separate sections being joined 
by moulding resin around the prefabricated inter- 

Fig 9: Front end structure in hybrid metal/plastic 

Fig 10: Metal/plastic hybrid door structure 

Fig 7 : Mg and Al used in seat frame 



locking points or by means of clinching integrated 
into the mould. It is said that for recycling purposes, 
it takes only a few seconds to break a metal/plastic 
composite, using a hammer mill, and that the resin 
element has properties akin to the virgin material, on 

Sample applications include instrument panel 
supports. In the cross-car beam co-developed by 
GM’s Delphi a40% reduction in weight was obtained 
against the replaced assembly, along with a 10% 

reduction in component and investment cost, for the 
same performance capability. Car front-ends with 
integrated bumper systems are further good candi- 
dates, Fig 9. A research project has also been carried 
out on car side doors having sufficient structural 
integrity to transmit impact forces from A- to B-post 
in the closed position. Fig 10 shows a sample door, 
requiring no further framework to support wing 
mirror, lock or other door ‘furniture’. Seat frames 
with integral belt-anchorages have also been made 



for the Mercedes-Benz Viano minibus in the process. 
The first volume car to incorporate the technology is 
the Audi A 6 . The front-end design was developed in 
association with the French ECLA company. The part 
is injection moulded in one piece and incorporates 
engine mountings, together with support for radiator 
and headlights. 

Bayer have also been active in a joint-venture 
glazing project with GE Plastics. A $40 million 
development programme is under way on abrasion- 
resistant, coated polycarbonate vehicle windows as 
an alternative to glass systems. An abrasion-resist- 
ance equal to that of glass is targeted and high- 
volume production of windows is foreseen. Some 
40% weight saving is forecast against equivalent 
glass systems and greater design freedom plus higher 
impact resistance is also claimed. Better resistance to 
forced entry, and reduced injuries in side-impact and 
roll-over accidents, have also been mooted. 

Structural design 
in polymer composites 

Work reported from Delphi Interior and Lighting 4 has f,g 73 • Audi A6 sun-roof 
shown the possibility of constructing instrument 
panel crossbeam and display panel assemblies in 
injection-moulded plastics. The structural crossbeam 
provides both structural integrity, to the assembly, 
and an air passage for the primary air vents; it 
attaches to the A-pillars at each of its ends and 
operates as a simply supported beam. In one concept 
by the company, Fig 11, the beam supports the 
display panel but is not integral with it. In the second 
concept, Fig 12, the panel and beam form an integral 
structure. The advantage of this second concept is the 
allowance of further integration of components into 
the single total assembly. 

Both concepts were built for retrofitting into a 



the initial objective was to develop a design for the 
plastic structures that would provide static stiffness 
equivalent to that of the baseline steel structure. The 
first natural frequency of the beam was also matched 
to the steel baseline. Both objectives were met with- 
out resort to wall thicknesses that would have com- 
promised conventional injection-mould cycle times. 

The next requirement was to meet FMVSS 
impact energy absorption levels resulting from input 
loads to the knee bolsters supported by the system in 
a 30 mph crash. It was found that materials with strain 
rates as low as 2% could absorb this energy without 
overload to the occupant’s femur. Finally the struc- 
ture had to meet these performance levels under 
specified temperature and vibrational environments, 
necessitating an evaluation of creep and fatigue 
capability. Temperature range of -30 to +190 F and 
fatigue criteria were based on subjection to service 
loads measured under real road conditions. 

Retention of properties USCAR/EWCAP Class 3 
Temperature/Humidity Testing 

Tensile strength, Mpa 

1 5% GR PBT 



30% GR PBT 



(30% Glass) 







After 40 cycles* 






Retention, % 






Elongation, % 







After 40 cycles* 






Retention, % 






• Each cycle consists of 6 hr. at 90'C/95% Rh followed by 2 Hr. at 1 25'C with humidity vented 

Recycled PET, and prime PBT, 
for sun-roof parts 

Following the recent interest shown in Chrysler’ s 
plan forusing low-cost, glass-reinforced polyethylene 
terephthalate (PET) for its China car body construc- 
tion, comes an announcement from DuPont that it is 
supplying recycled PET for comer mouldings of the 
Webasto sun-roof of the Audi A6, Fig 13. The 
mouldings hold the sliding roof frame’s aluminium 
rails together and incorporate guides for the cable 
that moves the roof and channels off rain-water. The 
material has withstood 30,000 open-close cycles in 
trials of the sun-roof. In another sun-roof application, 
DuPont’s polybutylene terephthalate (PBT) resin by 
Westmont Technik of Dusseldorf is used in a new 
design of sliding roof incorporating a built-in Venetian 
blind, Fig 14. The PBT slats of the blind can be 
adjusted to deflect cooling air into the vehicle in hot 
weather. The company has also introduced a PBT 
grade, Crastin 5000, for electrical connectors and 
parts. This has improved hydrolysis resistance for 
high temperature and high humidity operation. The 
first variants produced, for connectors, have 30% 
glass reinforcement and properties are seen in the 
table, left. 

Further details have also been released on the 
Chrysler China car which is now called the Compos- 
ite Concept Vehicle. Its shell structure is made up of 
four major elements, Fig 15, to fit over a steel chassis. 
Target cost of the PET-based GRP is less than £2 per 
kg, compared with £6.50 -£8.50 for SMC and SRIM. 







«ncuioes«c,wniocf.i fig ]2; Plastic structural instrument panel 



When the four shell elements are bonded to- 
gether enough torsional stiffness is available without 
the need for further reinforcement. Ashland Pliogrip 
fast-cure adhesive is used and cycle time for making 
the four main mouldings is three minutes, seven 
minutes faster than for equivalent sized SRIM mould- 
ings. The low cycle time has been achieved by a 
sequence of gate openings and gas injections into the 
mould to help guide the process, which was devised 
using computer-aided mould-flow techniques. An 
8500 tonne machine injects the PET resins into the 
145 tonne moulds, which each measure a massive 4.5 
X 2.5 X 2 metres. 

fibre has body trim potential 

Nissan has announced the development of what is 
claimed to be the world’s first structural-coloured 
fibre, in a joint venture with Teijin Limited and 
Tanaka Kikinzoku KK, Fig 16. The colour of the 
fibre is produced by its structure and does not involve 
the use of dyes. The principles of this technology are 
based on research into biometrics, an engineering 
practice that looks at prominent functions of living 
things and applies them in practical applications. 
One of the findings of this research is the principle 
behind the iridescence of the wings of Morpho butter- 
flies native to South America. Research has been 
conducted to apply this knowledge to the production 
of coloured fibres for interior and exterior trim mate- 
rials, including seat fabric. The structural-coloured 

fibre has been developed by applying the theory of 
multilayered thin-film interference to a cross section 
of the fibre. This is the principle behind the iridescent 
scales of Morpho butterfly wings. The fibre displays 
a high metallic sheen and a clear colour shade under 
exposure to different types of light, including natural, 
fluorescent and incandescent. The colour properties 
and tint also vary depending on the light source used 
and the viewing angle. As the fibre does not use any 
dyes, the production of the fibre reduces environ- 
mental pollution due to waste dye solutions. It also 
prevents any worries about skin rashes or other 
allergic reactions to dyes. 

Fig 15: Composite Concept Vehicle shell 

Fig 16: a: right, 
Reflective spectra of 
some samples; b, 
opposite, the 
technology explained 



Kefkcliw spevlru of iht muliil.m i model 



Material property charts and 
performance indices 

A rigorous method for evaluating the choice of 
materials for new designs has been made by a re- 
searcher 5 of the Cambridge University Engineering 
Design Centre which avoids the ‘hunch’ or ‘guessti- 
mating’ approach so common in making these deci- 

The author points out that it is seldom that the 
performance of a component depends on just one 
property of its constituent material. So by plotting 
one property against another (stiffness and density, 
say, for lightweight components), and mapping out 
the fields of materials in each material class, rapid 
assessment of material suitability is possible. 

Fig 17 shows how, on a log scale plot of elastic 
modulus and density, the data for a particular class of 
materials clusters together in envelopes. The sloping 
lines show how other derived properties, based on the 
two basic ones, can also be plotted on the graph. The 
elongated envelopes refer to material classes in which 
there is structure-sensitivity, in that properties can be 
altered significantly by such techniques as micro- 
alloying and heat treatment. While the modulus of a 
solid is a relatively well defined quantity, the author 
explains, strength is not and the envelopes take up 
different shapes. Fig 18. The shapes alter again when 
fracture toughness and other properties are plotted 
against density and the author provides explanations 
for these in the book, based on the behaviour of the 
material’s molecular structure. 

In general the charts show the range of any 
given property accessible to the designer and identify 
the material class associated with segments of that 

Logical design involves asking such questions 
as ‘what does the component do; what is to be 
maximised and minimised; what non-negotiable con- 
ditions must be met and what negotiable and desir- 
able conditions . . .’, defining function, objective and 
constraints. Property limits can be derived from these 
together with indices that will optimize material 
selection. Performance indices are groupings of the 
material properties which can maximize some per- 
formance aspect of the component, ratio of modulus 
to density being the classic one of specific stiffness, 
which would be a key consideration in the design of 
a light, stiff tie-rod. Many such indices exist each 

characterizing a particular combination of function, 
objective and constraint, as in the beam case, Fig 19. 

In this ‘index and chart’ material selection 
procedure, the index isolates the combination of 
material properties and shape information that 
maximiszes performance prior to using the chart for 
selecting the appropriate material as described above. 
Fig 20 shows how the index derived by isolating the 
key parameters from the equations describes the 
objective function after the constraints have elimi- 
nated the free variables. 

Design of a mechanical component, Ashby 
asserts, is specified by three groups of variables; the 
functional requirements F (need to carry loads, trans- 
mit heat, etc.), geometric specifications G and some 
combination of properties p so that maximized per- 

P =f(F,G, p) 

When these three groups are separable such that 

P = f,(F), f 2 (G), f/p), 

as they usually are, the optimum choice of material 
can then become independent of many of the design 
details and a merit index can be much more simply 
defined. Thus the steps in Fig 20 involve identifying 
primary function, writing down the equation for the 
objective (could be the energy stored per unit volume 
in a spring example) then identifying the constraints 
that limit the optimization by this last process in 
terms of a specific value of stiffness, say. The free 
variables can then be eliminated in leading to the 
index value (the objective function). 

In the case of a beam, for a simply supported 
length L, under transverse load F, there might be a 
constraint on the stiffness S such that it must not 
deflect more than 6 at mid-span. Here 

S = F/8 >/= CEI/L 3 

where E is modulus and I section second moment of 
area, C being a constant depending on load distribu- 
tion. For a square section beam, width b, I is A 772 
whereA is section area , which can be reduced in order 
to lessen weight, but only until the stiffness constraint 
is met. Eliminating A in the objective function gives 
the mass 



m >/= (C12S) I/2 (l 5/2 )( p/E ,/2 ) 

and the best materials for a light, stiff beam are those 
with large values of the material index 

M = E m /p 

where p is density. 

Because section shape is critical to bending 
stiffness a shape factor is defined as 

0 - 4nl/A 2 

which is dimensionless and can take values as high as 
100 for very efficient sections. 

Cross-substitution of these equations can be 
carried out to eliminate the free variables and a 


Fig 19: Identifying the primary function 

Density (p). Mg/m 3 

Fig 17: Specific stiffness property chart 



modified performance index would include a shape- 

Fig 18: Specific strength property chart 

Density(p), Mg/m 3 



Design for self-pierce riveting 

More widespread use of aluminium alloys, and light- 
weight coated steels, in automotive body construc- 
tion is leading to alternatives to the old certainty of 
spot- welding. Textron’s Fastriv system is proving a 
reliable substitute, a leak-proof system with fully 
automated installation equipment, and claimed greater 
dynamic strength than spot- welding. 

The Fastriv self-piercing fastening system, Fig 
21, can join two or three metal sheets with combined 
thickness from 1.5 to 8.5 mm. In the placing sequence 
of Fig 22, a tubular rivet is driven by a hydraulically 
operated press tool into the materials to be joined; 
first the rivet pierces the top sheets then it flares 
radially into the lower sheet against the die, without 
breaking through, and forms a mechanical interlock, 
the assembly process being complete within 1-2 

A single Fastriv joint is said to be close to the 
shear strength of an equivalent spot- weld and stronger 
in peal strength, without the disadvantage of a heat- 
affected zone; the suppliers say at least 30% of spot- 
welding applications could be displaced by its use. 
There must, however, be access for tooling on both 
sides of the sheets and sheet thickness/hardness must 
fall within a given range. 

Optimized Fastriv joint performance is ob- 
tained when the sheet material with the higher elon- 
gation is placed on the die side, when joining sheets 
of dissimilar material. For joining sheets of dissimi- 
lar thickness, the thinner sheet should be placed on 
the punch side and a large-head rivet should be used. 
Since joint elements of self-pierce fastening are not 
symmetrical on both sides, joint arrangements need 
special attention and suggestions are given in Fig 23. 

Suggested automotive application areas are 
shown in Fig 24, the company recommending com- 
bination with adhesive bonding for load-bearing 
carrier beams. Steel panels can be joined to alu- 
minium ones, where necessary, and the pre-painted 
panels can also be successfully joined by the process. 
A number of finishes are available on the rivets to 
ensure against corrosion. The standard finish used is 
electroplated bright zinc but Almac mechanically 
applied coating has higher corrosion resistance than 
zinc. Kalgard has an organic binder containing alu- 
minium; Deltaseal has an organic microlayer top- 
coat with PTFE; finally, Dacromet has an organic 

binder containing both zinc and other protective 

Similar to that for spot-welding, installation 
equipment can be floor-mounted or portable, suspen- 
sion-mounted, and have manual or robotic manipula- 
tion. Fasteners can be either tape-fed from reels or 
blow-fed from a vibrating bowl, through plastic 
tubing. Fig 25 shows proportions of the riveting 
module, from which an idea of joint access can be 
obtained. Various rivet head forms are available as 
well as the standard oval head, including flat counter- 

Fig 21: Fastriv configuration 

Fig 22: Installation 

Fig23: Suggested rivet 



sunk and a ‘tinman’ head which is a flat head which 
stands proud for occasions when a visual feature is 
required to be made. Head diameters can range from 
5.38 to 9.78 mm for rivet diameters from 3.10 to 4.78 
mm. For the 5 mm (nominal) rivet size, fatigue 

1,000 10,000 100,000 1 , 000,000 

Life Reversals 

Sourca: Rovar Group lid 

Fig 24: Automotive body applications (below) 
Fig 25: Riveting module sizes (right) 

Fig 26: Fatigue performance (above) 





SIZES -D' (mm) 





A. 25 













"G" (mm) 







1 00 













Chapter 2: 

Structure and safety 

Vehicle structural design innovation continues to be driven 
by the impetus for improved passenger protection and light- 
weighting for lower fuel consumption and reduced C0 9 
emission. With the perfection of controlled collapse front- 
ends the next stage in development is an analysis of footwell 
deformation so that controlled-collapse can be extended to 
that region. For further integrity of the occupant survival shell 
attention is also being paid to pillar/rail joint stiffness in 
reducing the likelihood of body shell collapse. Techniques of 
structural analysis are also being used for the prediction of 
noise level within the body shell. Attention is also finally 
turning from the protection of the vehicle occupant to that of 
the pedestrian road-user and realistic pedestrian models are 
now being produced for authentic laboratory crash analysis. 
In vehicle packaging, generally, is beginning to focus on the 
ageing and infirm section of the population and occupant 
restraint systems are being designed to be adaptive in their 



Understanding footwell 
deformation in vehicle impact 

Recently published work by researchers at the US 
Automobile Safety Laboratory, and its collabora- 
tors 1 , is suggesting a way of better understanding the 
collapse mechanism for a passenger-car footwell 
during frontal and front/side impacts. Hitherto, the 
authors maintain, the study of structural intrusion in 
this region has been based on toe-board mounted 
accelerometers set up to measure in the direction of 
the vehicles longitudinal axis. Since the instrument 
itself can easily be disoriented from its measurement 
axis, during the intrusion process, reliable dynamic 
deformation data has been difficult to forecast. In 
particular, a three-dimensional characterization of 
the collapse and buckling of the thin metal structures 
is difficult to obtain, except perhaps using X-ray 

speed-photography techniques for a qualitative ap- 
praisal. Obtaining orthogonal frames of reference is 
considered difficult with any conventional cine-photo 

Post-crash measurements can only provide in- 
trusion data for the last instant of impact and no 
elastic effects occurring during or after the crash can 
be evaluated — nor can a progressive sequence of 
toe-pan translations and rotations be obtained. Hence 
the stimulus for these researchers to use a sled system 
for obtaining a toe-pan deformation pulse, akin to 
that obtained by sleds for vehicle deceleration as a 
whole. To design an intrusion system using a labora- 
tory sled, work commences with obtaining a simple 
translational intrusion profile parallel to the ground. 
Fig 1 . The intrusion simulator has been designed to 
allow rotational components and asymmetrical load- 

Fig 1: Post-crash profile of 
toe-pan intrusion 



80 - 

60 - 
50 • - 
30 - 
20 -- 
10 -- 
0 — - 

Offset rigid 

Intrusion Simulator 
Final Toepan 


10 20 30 40 50 60 

cm (From arbitrary reference location) 

Fig 2: Hydraulic 
projected view of 
programmed row 
of holes 


I*- 7.62 cm 

o o o 

OOOOOOOOO- outpu! Row 





Main Portion 

Tibia Axial Load (N) Deceleration (g's) 






Guide Rail - 

Fig 3: Intrusion simulation mechanism 


Fig 4: Sled deceleration vs toe-pan acceleration in 
a sled frame 

20 40 60 80 100 120 140 160 

Time |m»| 

Fig 5: Left distal tibia load for intrusion and no- 
intrusion sled tests 

ing effects, at a later stage. 

The toe-pan produces a typical 20 kN force 
applied to both legs, about twice that of the axial 
tibial breaking strength. The panel is generally capa- 
ble of 25 cm maximum displacement during a crash 
event and it has been found by accident-studies that 
85.6% of below-knee injuries are sustained with less 
than 14 cm intrusion. The sled therefore moves the 
toe-panel in a progammed displacement time profile, 
with these maxima, by integrating the actuation of 
the panel components with the function of the sled’s 
hydraulic decelerator. The sled and test buck are 
typically accelerated to a predetermined velocity 
and decelerated according tot he crash pulse. The 
decelerator is configured as a cylinder with an array 
of discrete orifices over its length. Fig 2 suggesting a 
flat-pattern view. At the start of an impact, hydraulic 
pressure is transmitted to both sides of the slave 
cylinder, holding the toe-pan stationary with respect 
to the buck. When the decelerator piston passes the 
control-side output port, pressure on that side drops 
and the toe-pan moves under the force of the hydrau- 
lic drive lines remaining in front of the sled piston. A 
constant force internal stop in the cylinders at the 
maximum mechanical stroke is used to determine 
maximum toe-pan travel according cylinder posi- 
tion. The intrusion pulse is controlled by variation of 
orifice sizes on the drive and control sides. Fig 3 
shows the physical set-up, the simulator consisting of 
a toe-pan carriage with separate footrests for each 
foot. Toe-pan carriage travel is 25 cms and the 
hydraulic cylinder diameter is 6.4 cms. The rear 
mounting position is adjustable to limit toe-pan 
travel in increments of 2.5 cm. Controllable toe-pan 
intrusion parameters, during the impact event, in- 
clude initiation, by location of a control line within 
the hydraulic decelerator; stroke, controlled by cyl- 
inder mount location and cylinder stop; and accelera- 
tion profile, depending on a combination of passive 
orifices in the drive lines forming an hydraulic accu- 
mulator so that dynamic effects of the sled pulse can 
be used to tailor the toe-pan pulse. 

The researchers demonstrated the use of the 
intrusion simulator with sled tests carried out with 
zero to 7.9 cm intrusion, using the ALEX dummy 
developed by NHTSA. The test buck was configured 
as a mid-size car and based on a crash pulse measured 
by damped accelerometers. Sled velocity at initia- 
tion of the crash pulse was fixed at 58.2 km/hand full 



intrusion stroke is selected to be at 7 1 ms into the sled 
impact. In Fig 4 of a representative sled decel pulse 
and toe-pan intrusion pulse, intrusion decel relative 
to the test sled is approximately sinusoidal, with 80 
g peak, or 100 g relative to the ground. Net effect of 
sled and toe-pan decels is movement of the toe-pan 
into the occupant compartment, providing a poten- 
tially realistic translation intrusion pulse most rel- 
evant for lower extremity injury. Left tibia axial 
compressions are shown in Fig 5 for both representa- 
tive and no intrusion cases, the significant effect of 
toe-pan intrusion being demonstrated after about 70 
ms. The second peak of the intrusion means tribial 
axial compression is more than doubled from the no- 
intrusion run. 

roof rail roof rail 

Pillar to rail joint stiffness 

Low frequency vibration characteristics of a vehicle 
body structure are considerably influenced by the 
compliance of joints between the main pillars and 
rails, according to researchers at Kookmin Univer- 
sity 2 . While these effects have been examined at the 
design stage by using torsional spring elements to 
simulate the joint stiffnesses, apart from the inaccu- 
racy of the assumption it has been difficult to analyse 
the whole structure using the NASTRAN optimiza- 
tion routine because these mass-less spring elements 
cannot be represented. Because the rotational stiff- 
ness of the joint structure has a major influence, the 
authors have proposed a new modelling technique for 
the joint sub-structure involving an equivalent beam 
element. Fig 6 shows the traditional and proposed 
modelling techniques applied to a B-pillar to roof 
joint. The joint stiffness is computed using familiar 
finite element methods, Fig 7. Unit moments are 
applied at the tip of the pillar, with both ends of the 
roof-rail fixed. FE static analysis gives the deformed 
rotation angle from which rotational stiffness can be 

Fig 6: Traditional and 
proposed modelling 

f | initial 
m optimal 



In the subsequent MSC/NASTRAN optimiza- 
tion run, if the section moments of inertia are chosen 
as design variables, their optimal values can be 
obtained to meet the vehicle’s frequency require- 
ments. The table in Fig 8 shows a design change 
guideline to obtain optimum joint stiffness in this 
respect. The joint stiffnesses in the X and Y planes, 
plus the rotational stiffness about the Z-axis of the 
section centroid can be obtained. When these results 
are plotted out as a bar-graph as in Fig 9, the authors 
explain that it is relatively simple to which stiffnesses 
need increasing and which could perhaps be reduced. 
The example in the figure implies that the Y-direc- 

tion rotational stiffness of the B-pillar joint needs to 
be increased while the other directional stiffnesses 
may be decreased. Detail study then showed the 
thickness and shape of the reinforcement panel needed 
to be altered in this case, establishing thickness and 
nodal co-ordinates of the panel as design parameters. 

The reinforcement may be optimally posi- 
tioned anywhere between the inner and outer panels 
of the joint, to satisfy the stiffness requirements, and 
by setting the objective function of the optimization 
run for minimizing mass, the constraints can be 
evaluated for achieving the required rotational 
stiffnesses, Fig 10. 

unit : % 




A-pillar to roof 




B-pillar to roof 



+7 0 

C-pillar to roof 



FBHP to A-pillar 



-16 1 

FBHP to rocker 




B-pillar to rocker 




C-pillar to QLP 




Fig 7: FE model of joint 

Fig 8: Design change guideline 





+ 34 15 


+ 7.18 



+33 33 



refer to Fig 8 

Fig 10: Joint configuration 



Structure analysis for interior noise 

A researcher at Rover 3 insists that refinement at his 
company is now designed-in at the concept phase of 
new products. In a recent paper he differentiated 
between component and system modelling for NVH 
and pointed out the difficulties of examining high 
frequency behaviour, above 70 Hz. At these fre- 
quency levels, the author explains that vehicle prob- 
lems are dominated by flexible modes of the compo- 
nents direct-connected to the body and care is needed 
in interrogating information from the standpoints of 
non-linearities arising, for example, in rubber an- 
chorages and structural complexity where crude beam 
models are typically 25-30% inaccurate. 

Suppliers are now requested to carry out de- 
tailed modelling for analysis purposes with predic- 
tions of natural frequency for something like a 
subframe being to 10 % inaccuracy. Fig 1 1 compares 
a beam-element model with an accurate supplier 
model for a subframe and Fig 12 shows the driving 
point modal response of these two models. While it 
can be seen that the mass is predicted well in both 
model cases (amplitude at low frequency before 
resonance is proportional to the mass of the compo- 
nent), the bem element model only recognises two 
main resonances and misses the smaller ones be- 
tween 300 and 350 Hz because these modes are out 
of the plane of excitation they are poorly excited. 

Above 300 Hz, the author argues, FE model- 
ling is unsuitable because the non-linear behaviour of 

Fig 11: Beam model vs superelement 

most systems gives poor levels of prediction and 
experimental test methods have to be used. Cur- 
rently, statistical energy analysis (SEA) techniques 
are being evaluated in conjunction with ISVR and a 
number of trim suppliers that eventually will allow 
accurate prediction of vehicle interior noise from 300 
Hz upwards. Rigid lumped-mass representations of 
bodies are giving way to models which allow the 
effect of flexible modes. However, usually structural 
models for crashworthiness analysis are unsuitable as 
they do not incorporate closures and trim. Since the 
BMW acquisition of Rover the ability to add trim 
items to the models has been developed. Large trim 
areas such as head-linings are modelled as non- 
structural masses, additional damping effect being 
added to the elements representing the adjacent 
panel. For other items such as facia panels, which add 
stiffness to the structure, tuned spring/damper ele- 
ments are added to the structural model. 

The author suggests that a current 90,000 ele- 
ment model for a vehicle body will increase in terms 
of number of elements by 50% every time that a new 
model is introduced. He thinks model solution time 
will remain constant as machine speed is increasing 
at similar rate. Another development is studying 
effects of modifications to body structures by means 
of super-element whole-vehicle models which com- 
bine body and powertrain. Fig 13. The company also 
employs SYSNOISE software for acoustic model- 
ling activities, specifically for problems with engine 
intake and exhaust systems. Vehicle interior acoustic 
modelling, Fig 14, has been used for some time to 
assess panel sensitivity and work is underway to 
couple the acoustic models to the full-vehicle mod- 

A case study in assessing the refinement ben- 

Force Through Mount (arbitary unitit 



Fig 13: “ Whole vehicle” model 

Fig 14: Acoustic model of vehicle cabin 

Fig 16: FE model prediction 

efits of a compliantly mounted front subframe was 
given by the author. Fig 15 shows typical interior 
noise measurement for a vehicle driven at constant 50 
kph along a surface of 20 mm cold-rolled chippings. 
At 25-45 Hz global modes of the vehicle body 
structure are evident; from 75-130 Hz excitation of 
the 1st and 2nd longitudinal cavity modes of the 
vehicle occurs and main noise source is tyre tread 
blocks impacting and releasing from the road sur- 
face; at 220-260 Hz strong excitation is generated at 
the wheel hub by an acoustic resonance of the air 
within the tyre, seen as two modes describing waves 
travelling forwards and backwards within the tyre. 

The proposed subframe involved a front sus- 
pension system with body mounted MacPherson 
strut with an L-shaped lower arm mounted to the 
subframe at two locations; steering rack and anti-roll 
bar are similarly subframe mounted. The lower tie for 
the engine mounting is also attached to the subframe 
but the upper tie-bar is body-mounted. There is also 
a primary vibration route via the drive-shafts to the 
wheel hubs and back through the suspension system, 
the primary one for structure-borne high frequency 
gear noise. 

When in-service recorded load inputs were 
applied to the model, results for the left hand front 
subframe mount are shown in Fig 16 as a level of 
force going into the body (not interior noise). Below 
1 00 Hz there are additional resonances attributable to 
rigid body modes of the subframe on its mounts. 
Beyond 100 Hz there is marked improvement for the 
compliantly mounted set-up. 



According to a researcher of Ove Arup and 
Partners, in his paper to the recent IMechE Vehicle 
Noise and Vibration conference 4 , NVH refinement 
rivals crashworthiness as the major contemporary 
automotive design target. The author described an 
indirect boundary element method used to predict the 
noise sensitivity of a vehicle body. A modal model of 
the vehicle body is coupled to a boundary element 
model of the interior cavity, with unit forces, in the 
three transational degrees of freedom, applied at the 
engine mounts and damper towers. The method then 
depends on noise transfer functions being predicted 
between these force input locations and the occu- 
pants ear positions. 

In normal finite element modelling, the author 
explains, predicting the body acoustic sensitivity 
starts with generating a finite element mesh, which 
might even simulate glazing, closures and trim (it 
could contain 200,000 elements for extracting some 
800 modes); normal modes in the frequency range 
below 200 Hz are then extracted to obtain vectorial 
data, along the modal damping ratios, for making the 
modal model. 

The airspace cavity is modelled using 10-20 
modes, depending on cabin space geometry. A modal 
damping model is used which effectively ‘smears’ 
the damping over the entire model rather than just 
representing local area absorptions such as seat 
surfaces. Wavelengths of these acoustic waves are 
much greater than those of the body structure in this 
under 200 Hz range so elements up to 50-100 mm in 
length can be used for the model compared with 
element lengths for the structural model of 10-20 

The two models are next coupled together and 
the loads applied to solve for airspace pressure. As 
the design progresses the models can be modified to 
account for the greater complexity brought about bey 
added features. At the early crude modelling stage 
focus is on the low-frequency noise sensitivity. The 
main drawback with FEM, however, is the time 
required for mesh generation. 

With boundary element modelling (BEM), is 
suitable for acoustics applications where there are 
exterior or interior fields. The indirect BEM is pro- 
posed by the author in this case because the interior 
noise problem has several domains, cabin, seats and 
boot airspace. While BEM uses the same structural 
modal model as the FEM method described above, 

but cabin and boot airspaces are modelled using 2-D 

A typical structural model is in Fig 17 and a 
boundary element model in Fig 18. The latter can be 
created in a proprietary pre-processor and can be 
rapidly created from the structural mesh. A field 
point mesh, Fig 19, has also to be generated so 
occupant ear sound pressure level (SPL) functions 
can be calculated and the pressure field visualised. A 
surface impedance model is used to simulate seat and 
other trim absorption. The indirect BEM model 
works in the inverse of impedance, the admittance 
(velocity normal to the element divided by pressure). 

The coupled indirect BEM requires a so-called 
surrogate structural model (SSM) and an acoustic 
model. The structural modal model is transferred on 
to the SSM and the search algorithm maps a vector 
component of each of the structural nodes on to the 
nodes of the BEM model. The modal model is 
considerably reduced by this process as the number 
of structural nodes is much greater than those of the 
BEM. For the example in the figures, a 2000 element 
acoustic model is involved; there are 531 structural 
modes, 91 frequencies and 24 load cases so consider- 
able computer power is involved. 

Results of the example analysis illustrated here, 
solved by supercomputer, gives the NTFs displayed 

Fig 18: Cabin acoustic model 

Fig 19: Cabin field point mesh 



on a magnitude/phase plot, Fig 20, and cabin/boot 
airspace sound pressure distribution as in Fig 21. 
Where NTFs are above target, as at 144 Hz a noise 

Fig 17: Vehicle body FEM 

boom occurs. Panel contribution and modal partici- 
pation analysis is used to determine the controlling 
NTF factors at this frequency. 

Fig 20 left; Vehicle body NTF, phase pressure top and NTF 
magnification (dB/N) pressure bottom 
Fig 21 above; Interior pressure distribution 



Preparing for statutory 
pedestrian protection 

The next stage in structural design for safety, of 
trying to design-in protection for pedestrians subject 
to vehicle impacts, is being addressed with vigour by 
the UK Transport Research Laboratory 5 whose first 
priority is developing pedestrian impactor physical- 
models, to be a base for the development of standard- 
ised testing. 

Since the historic work of Ralph Nader in the 
1960s, culminating in his best-selling book ‘Unsafe 
at any speed’, little has been done for that most 
vulnerable of roadway cas ualties, the pedestrian. 
Even if motorists can be persuaded to respect the 
speed limits in intensively pedestrianised areas, or 
have his/her speed more forcibly reduced by traffic- 
calming measures, much can be done to vehicle 
front-end structures to increase the chance of pedes- 
trian survival under impact. 

According to TRL researchers a recent survey 
has shown more than one third of road deaths were of 
unprotected road users such as pedestrians and pedal 
cycles, some 60% of whom are struck by car front- 
ends. The laboratory produced an experimental safety 
car in 1985, with front-end modified using conven- 
tional materials, which demonstrated a considerable 
improvement in reduced injury level in 40 kph im- 
pacts, but this was largely ignored by UK legislators, 
despite DoT funding, in terms of ruling in measures 
to reduce the death/serious-injury toll of up to 100,000 
per year. 

Key impact areas, in order of occurrence after 
first strike, are lower-limb to bumper, femur/pelvis to 
bonnet leading-edge, and head to bonnet-top. There- 
fore, in response an EC directive, arising from re- 
search by the TNO-chaired Working Group, impactors 
now being developed represent knee-joint shear force 
and bending moment; femur bending moment and 
pelvis loads; child and adult head accelerations. Each 
impactor is propelled into the car and output from 
associated instrumentation establishes whether en- 
ergy absorbing characteristics of the car are accept- 
able. TRL’s responsibility is development of the 

Initial impact with pedestrians is confirmed to 
be mostly side-on, in road-crossing situations; the 
bumper contacts typically below the knee and accel- 
erates the lower leg against the inertia of the body 

The shear displacement (d) can be found from the following formula; 

d - e x sin D 

upper leg-form to bonnet leading-edge test method 
and the laboratory now supply both leg-form and 
upper legform impactors to the car industry. They 
overcome the unreliability of testing using conven- 
tional dummies. 

Fig 22a: TRL knee joint at maximum bending 
displacement, above, and maximum shear 
displacement, below, showing trigonometric 
calculations for actual displacements 



Fig 22c: TRL leg-form impactor 

above and below the contact point, reaction forces 
causing bending and shearing with reactions between 
foot and road, hip-joint and upper body mass, adding 
to these forces. Resulting injuries are bending and 
local crushing of leg bones and/or knee ligament 
injury through bending and shearing. 

The legform impactor thus consists of ‘femur’ 
and ‘tibia’ sections joined by a mechanical knee, with 
sideways loaded joint stiffness, the elements of the 
impactor having physical properties akin to a 50th 
percentile male. Studies with cadavers, and compu- 
ter-simulations, have shown that leg inertia forces 
have major influence but foot-road reaction and 
upper body inertia only minor influence on these leg 
forces. The impactor is not ‘fooled’ by palliative 
efforts to soften only the bumper area which might 
reduce lower leg injury at the expense of knee joint 
damage. The TRL ‘knee’ comprises an elastic spring 
to produce shear stiffness and disposable deformable 
ligaments for the bending stiffness. The elastic spring 
is in the form of a parallelogram, a pair of bending 
ligaments joining the knee ends of each leg element; 
this particular configuration provides necessary shear 
stiffness whilst being unaffected by the bending 
moments required to deform the bending ligaments. 
Individual potentiometers separately determine bend- 
ing and shear displacements, their angular outputs 
being converted trigonometrically, Fig 22. 

The design of the upper legform impactor is 
geared to the realisation that while first contact is 
made between lower leg and bumper, next phase is 
for the pedestrian's legs to be swept away from 
beneath with contact then occurring between bonnet 
leading edge and upper leg, the first contact having 
been found to have an effect on the second one by 
modifying upper leg angle and impact velocity to an 
extent dependent on vehicle shape. Principally the 
upper leg impact causes bending of the femur and 
corresponding forces in the hip joint, the mass seen 
by the car bonnet being a combination of that of the 
upper leg and the inertia of body parts above and 
below it, its effective value again dependent on body 
shape. To determine the effect of shape, testing a 
wide range of body configurations has led to the 
recording of data in the form of look-up graphs for the 
use in the eventual assessment procedures. 

Fig 23 shows the current upper legform impactor 
which comprises vertical front member representing 
the adult femur, supported top and bottom by load 



transducers on a vertical rear member, in turn mounted 
on the end of the guidance system through a torque- 
limiting joint. The front member has strain gauges to 
measure bending and is covered with a 50 mm foam 
layer to simulate flesh. Since the test procedure has 
been designed to make the impactor aim at a line 
defined by the car’s shape, highest bending moment 
does not necessarily occur at the midpoint, if the cars 
peak structural stiffness occurs above or below it, so 
measurement is made at midpoint and 50 mm above 
and below. Provision is made to attach weights to the 
impactor so its mass can be adjusted to the effective 
mass of the upper leg on impact appropriate to each 
car shape, described by the look-up graphs. The 
torque-limiting joint is specified to protect the guid- 
ance system from damage on cars with poor impact 
protection, its minimum torque setting providing a 
rigid joint on cars meeting the desired protection 

To meet the knee shear and lower leg accelera- 
tion acceptance criteria the car front must be able to 
absorb energy whilst to meet knee-bending criteria 
the car must make a well distributed contact along the 
length of the legform. While the second criterion 
might imply a requirement for a vertical front-end, in 
fact the TRL authors say that local deformation 
effectively “converts a streamlined shape into a more 
upright one”. While the design of bumpers to meet 
mandatory low-speed impact legislation do provide 
some advantage to pedestrian protection, the re- 
searchers say it is those which have a substantial 
distance between the plastic bumper skins and the 
strong underlying structure that best meet the re- 
quirements of the legform test. Ideally a locally 
deformable lightweight outer bumper skin should be 
supported by an integrally stiff energy absorbing 
foam reacting against the strong underlying struc- 
ture, the crush depth being at least 40 mm, A deep 
bumper/spoiler making a low contact with the pedes- 
trian’s leg at the same time as the bumper would 
reduce bending cross the knee. 

In terms of upper legform impact, for a car with 
bonnet leading edge height of about 700 mm the 
depth of crush required would be 110 mm for a 
typical 100 mm bumper lead. At 900 mm bonnet 
leading edge height the crush depth requirement 
increases to 210 mm under the specified impact 
loading. The limit on bending moment also means 
that the contact force on the impactor must not be 

concentrated at one point, requiring generously con- 
toured front ends. Usually the existing heavy outer 
bonnet reinforcing frame should be weakened and 
moved slightly backwards so as to allow easier 
deformation of the leading edge. The bonnet lock 
should also be move back and deformable clear 
plastic headlamp front-faces should be employed and 
a revised mounting system allow the lamps beneath 
to collapse inwards into space that is normally avail- 
able rather than rigidly mounting them to a bulkhead. 

In terms of the head injury criteria, a limit on 
acceleration is the key requirement with both child 
and adult head-forms having a minimum stopping 
distance of 70 mm which reduces according to the 
amount from which the bonnet inclines from the 
normal-to-surface impact situation. Some current 
bonnet skins are acceptable provided underlying 
localised reinforcements could be replaced by more 
evenly spread systems. Proper clearance is also re- 
quired from suspension towers and engine surfaces, 
the authors point out. 




Design for the disabled 

According to researchers of the Royal College of 
Art 6 , a hard look at the realities of environmental 
stress, the ageing of populations and the crisis wel- 
fare systems are encountering, suggests that a new- 
concept “Mobility for all” vehicle may be a viable 
alternative to car ownership. At the same time, this 
shift from ‘A Car for All’ to ‘Mobility for All’ is quite 
in keeping with contemporary thinking about the 
shift from product to service and from the material to 

Fig 24: The “50+ coupe 

Fig 25: The “urban rickshaw 

the virtual. Environmentalists talk of reducing the 
global impact of human activity by a factor of be- 
tween five and 50 over the next 50 years, and a good 
point to start is transportation. The authors have been 
thinking about what this might mean and offer the 
following case studies of staff and student work as an 
insight into how this new mix of services and vehicle 
typologies might evolve. 

They argue that for the generation now moving 
into retirement, personal transport is synonymous 
with freedom, dignity and security. This suggests 
that mobility, in the future, will mean having a 
greater range of transport solutions at our disposal — 
a more integrated transport policy that blurs the 
traditional boundaries between public and private 
services and vehicles. Patterns of car ownership will 
change as more vehicles are banished from town and 
city centres, increasingly pedestrianised to ease con- 
gestion, improve security and provide a better quality 
of life for all those who work, live and pursue leisure 
activities there. While reservations about public trans- 
port — mixing with other people, lack of comfort and 
security, etc — can be greatly improved by interior 
layout, better seating, lighting to aid security and 
privacy, and higher grade interior materials and 
accessories to promote a feeling of ambient luxury. 

The increasing pressures to reduce materials 
and energy consumption in vehicle manufacture, and 
use, will lead to vehicles that last longer, can be 
adaptable throughout their working life and recycled 
or re-manufactured at the end of it. In simple terms 
cars will last longer and be more adaptable. For 
example, the traditional family car may encompass 
the needs of a growing family for its younger users, 
and act as a social space — and extension of the house 
— as they grow older. In America, older people are 
taking to the highways in their retirement to explore 
the vastness of their native country which they have 
been unable to enjoy during their working lives. 
Future trends indicate that vehicles and transport 
systems will have to respond to an increased empha- 
sis on freedom, versatility, variety of use and social 
interaction, and do so by exploiting technical ad- 
vances. Over the past twenty years personal security 
has become a prerequisite of everyday life: theft, 
vandalism and personal attacks are more prevalent 
and directed not specifically at the elderly or vulner- 
able. Consequently, security in and out of the vehicle 
has become a priority across the age ranges. 



Where restrictions in the range of movements 
are encountered, ergonomics remains an area of 
significant interest to drivers of all ages, in particular 
as expectations of comfort, fatigue reduction and 
personalisation of the driving environment have risen 
substantially. This leads to a great deal of interest in 
increased flexibility, adaptability and user-friendli- 
ness in the mainstream of vehicle design. 

The problem is that the major car manufactur- 
ers of the world continue to realise new designs 
around the slow and calculated evolution of a generic 
product. Experience tells them that if they get things 
wrong the costs can be enormous, like the profits if 
they get things right. So, advances in new directions 
are difficult to accommodate. Where age-friendly 
design can succeed is in offering a new direction for 
research and innovation, the results of which can be 
introduced to mainstream product ranges and tar- 
geted at the age sectors which are at present expand- 
ing and will continue to do so in future. Designed 
correctly, such improvements can appeal to a wider 
market sector by diminishing age differences and 
offering real benefits to younger purchasers — a ‘car 
for all’ that the marketing people can understand! 

The 50+ private car 

Vehicles targeted at particular user groups e.g. ‘la- 
dies’ or ‘oldies’ cars, can be seen as patronising and 
be rejected by both target group and other potential 
purchasers, whereas off-road and MPV’s have dem- 
onstrated the benefits of a large cabin area, easy entry 
and exit, high roofs and large door apertures which 
appeal to a majority of consumers and that can be 
particularly helpful for older people. The seat height 
allows the user to sit ‘on’ rather than ‘in’ the car, 
aiding parking, improving visibility, and making it 
easier to turn in the driving seat. 

Although few older people are wealthy, a sig- 
nificant proportion are ‘comfortably’ provided for 
and some have considerable disposable income. Two 
examples from Royal College of Art students Jim 
Das and Geoff Gardiner, address this market sector 
by exploring a ‘classic’ solution in terms of styling, 
taking care to integrate many of the age-friendly 
elements identified above to appeal to drivers of all 
ages. Features include sliding doors, level floor, 
spacious cabin, increased roof height, lowered boot 
access etc., and the interior components are designed 
to be adaptable: e.g. the seats can be upgraded to give 

extra support or to pivot for ease of entry. The driver 
is assisted with automatic gear change and power 
steering, and is offered joystick type controls in one 
design, which give the driver more space and make 
the car usable for people who would presently require 
an adapted vehicle. 

Instrumentation is by ‘head up display’, pro- 
jected on the windscreen to aid driver concentration. 
Information is ‘hierarchically driven’, and only what 
is important is presented at any time so as not to 
overload the driver. 

The format of a large coupe. Fig 24, reflects the 
lifestyle of a prosperous semi-retired couple, and the 
designs portrays qualities of sophistication and value. 
The cars features high shoulder lines to give a feeling 
of security and the front and rear sections have soft 
reactive surfaces that aid protection and reduce repair 
costs. The car would be manufactured from light- 
weight steels and re-cyclable plastics. 

Taxis and people movers 

The London Cab has been recently re-designed to 
offer more interior space and be wheelchair friendly. 
The London Cab is not only an icon of the capital, it 
also offers one of the most flexible transport systems 

Fig 27: Driverless taxi 



available. Ideally it should form the core of an 
integrated transport system within the capital, bridg- 
ing the gap between people’s homes and other trans- 
port services like the bus and underground. The 
problem is that it is relatively expensive and not easy 
to order for short trips, which means that for many 
older people it is not a viable travel option. A number 
of alternative scenarios have been investigated at the 
Royal College of Art, all of which seek to keep fares 
at an affordable level by reducing the capital and 
running costs of the taxi, use information technology 
to offer transport on ‘demand’ — with passengers 
booking by telephone or teletext — and extending its 
user profde to include disabled and older users. 

The urban rickshaw 

On London’ s congested streets a black cab is unlikely 
to travel more than 60 miles in a day, making an 
alternative power source (electricity, hydrogen, LPG) 
feasible. This would reduce space requirements and 
design limitations by eliminating engine and gear- 
box. Sotoris Kavros drew the inspiration for his 
lightweight taxi, Fig 25, from the Tuk-Tuk rickshaw 
of Indonesia, reworking the idea around an alterna- 
tive power source. The two passengers sit behind the 
driver, and the Urban Rickshaw has been designed to 
be as narrow as possible, to aid travelling through city 
streets and parking, while offering good headroom 
and easy access through doors that fold back along- 
side the vehicle. The exterior is styled to look ‘taxi- 
like’ with soft bumpers to prevent pedestrian injury. 

The greatly reduced capital cost of this vehicle 
compared with a new Metro cab costing £28,000, 
could cut fares in half and still offer the driver a good 

Short-haul taxi 

In designing this taxi, Fig 26, Peri Salvaag envisaged 
a future where large parts of the city are closed to cars. 
The function of this vehicle is move people from 
place to place within the city centre and link tourist 
and leisure attractions with public transport. The taxi 
runs on electricity, with the batteries forming the 
floor pan of the vehicle. It would be used for short 
journeys, typically lasting less than 20 minutes at 
relatively slow speeds, and would be able to enter 
pedestrianised areas. Safety is therefore an important 
feature, along with accessibility. The passengers 
would order the cab electronically — in the foyer of 
a museum say — and be advised about when to 
expect it to arrive. The driver is there to provide 
security and assistance, and payment would be by 
credit card. 

Short trips do not require the same level of 
passenger comfort as longer ones do and the vehicle 
exploits every inch of its tiny footprint with an 
interior design that features a centrally mounted 
driving seat with either two passengers perching 
behind on small half seats (bum rests), or a wheel- 
chair passenger who would enter through a door and 
ramp at the rear. The exterior is designed to look 
strong, stable and dignified. 

Driverless taxi 

The most radical proposal for a city taxi that has been 
investigated is for a driverless vehicle that would 
work within pedestrian zones. The taxi, Fig 27, 
follows a guidance system in the road surface and is 
in constant communication with a control station that 
locates vehicles as demanded, manages their jour- 

Fig 28: Kyoto tram 
( left and next page ) 



neys in the most efficient way and parks them in 
parking zones when not in use. During peak periods, 
the vehicles cruise the streets until ‘called’ by a 
passenger, just as in the case of the short-haul cab 
above. The individual vehicle then picks up the 
passenger, and the user also advises the drop-off 
point so that the control station can plan its next 

The cab is shared by up to four passengers to 
reduce costs and optimise use. It has large glass areas 
to give the passengers better visibility and a sense of 
security, while gull-wing doors aid entry and protect 
the interior from the weather. As there is no driver, 
the interior is designed so that passengers face one 
another around a central unit on which information 
about the city and route is projected. 

Shopping ferry 

This is a ferry service to be offered by a supermarket 
with the vehicle being commissioned or leased by the 
supermarket from the manufacturers. Out-of-town 
developments mean that more people rely on cars, to 
the detriment of older and disabled people. Back- 
ground research for this vehicle demonstrated that 
most supermarket journeys are weekly and less than 
one trolley-load of goods per person is purchased. 
Regular transport to and from appropriate points, or 
following a route around housing areas could reduce 
road traffic and build customer loyalty. 

In Mike Leadbetter’s scenario, the bus would 
be owned and run by the supermarket and would be 
called or booked by the customer, who would be 
collected from a point near their home. The size of the 

bus is optimised to work within conventional city 
streets and carries ten passengers and their shopping 
to keep the journey and waiting times down. Luggage 
provision is under each seat, wheelchair access being 
by ramped entry aided by external wheels which 
allow the bus to kneel, and the bus is driven by hub 
motors in the wheel rims, saving valuable space. The 
driver will provide security and assistance as re- 

Kyoto tram 

Kyoto is an historic city and the challenge here was 
to design a tram which suits a very wide user profile, 
including older and disabled people, mothers with 
children, and tourists, as well as commuters. The 
internal layout is spacious with groups of seats re- 
placing conventional benches, providing flexible 
seating options for single passengers, couples and 
people with small children as well as extending the 
clear floor area to aid wheelchair access. 

To further aid access the tram has a low floor 
and large doors that open flush with the sides of the 
vehicle and the interior has built-in grab handles. 

The tram, Fig 26, features large glazed areas 
that offer visibility to seated passengers and those in 
wheelchairs, as well as to standing passengers. Infor- 
mation is presented graphically to the passengers on 
the internal window surfaces with LCD technology, 
while the exterior of the tram has a low ride height to 
appear non-threatening and a soft, smooth front to 
minimise accidental injury. At night lighting effects 
similar to those used on aircraft would enhance the 
ambience, security, and safety of the vehicle. 



Adaptive restraint technologies 

Airbag systems is considered by Delphi Automotive 7 
to be the fastest growing sector of the automotive 
components market. 

The limitations of current technology centre 
around the fact that most airbag systems today deploy 
with constant force, regardless of occupant position, 
weight, size or seatbelt usage. Although the majority 
of airbag deployments occur in impacts with signifi- 
cantly lower speeds than are used for regulatory 
testing, current system designs make no distinction in 
the level of protection for restrained or unrestrained 
occupants, crash severity or the type or position of 
occupant involved in the crash. The next generation 
of airbag system will take these factors into account, 
monitoring occupants’ characteristics and the crash 
severity then tailoring airbag inflation to the situa- 
tion, Fig 27. 

Fundamental sensing and control elements for 
such systems include: characterisation of vehicle 
crash severity; detection and recognition of occupant 
size and location; sensing restraint system compo- 
nent configuration — seat belt usage, seat position 
and presence of a child seat; also adapting restraints 
to achieve maximum protection while minimising 
risk of injury. A Delphi team is developing sensors 
and restraints for Generation I ART systems targeted 
for late 1998 and Generation II systems targeted for 
2001, Fig 28. 

Crash severity sensing 

Current airbag systems utilise a Single Point Sensing 
and Diagnostics Module (SDM), in the passenger 

Fig 27: ASI passenger-side low-mount air-bag 

compartment, to detect a crash through vehicle de- 
celeration. Using a complex algorithm, the SDM 
determines the appropriate conditions and timing for 
deploying the airbag. Airbags are triggered at the 
lowest vehicle speed that might produce injury yet 
are designed to provide maximum restraint in the 
worst crash scenarios. Typically, airbags are de- 
ployed above 14-22 km/h Equivalent Barrier Speed 
(EBS) and require only 50-75 ms to deploy. Adaptive 
Restraints require an estimate of the potential occu- 
pant injury severity to make the best restraint activa- 
tion decision. Although they are not necessarily the 
same, crash severity can be related to occupant injury 
severity potential. Total change in vehicle velocity 
during the crash, for example, can be used to estimate 
the maximum occupant injury potential. This infor- 
mation is needed early in a crash, to determine 
severity and appropriate deployment. 

Due to the filtering effect of the vehicle struc- 
ture during a crash event, the total delta velocity may 
not always be established early enough with a single 
point sensor. To support the crash severity algorithm 
in the SDM, an accelerometer-based sensor located 
in the crush-zone of the car could be used. This 
information would be sent to the SDM and used for 
situations where multi-level deployments are avail- 
able. Delphi is also exploring Anticipatory Crash 
Sensing (ACS), which senses the speed at which the 
vehicle approaches exterior objects, to provide ear- 
lier crash severity estimates and significantly im- 
prove deployment decisions. A rollover sensing mod- 
ule is also being studied which can predict impending 
rollover and pitchover so as to trigger rollover safety 
devices, including seat belt reelers, inflatable cur- 
tains and side-airbags. Algorithm simulations sug- 
gest that it can predict rollover as much as 300 
milliseconds in advance of those events, critical for 
the occurrence of occupant injury. 

Occupant sensing 

The company is also developing technologies to 
measure occupant mass and position in order to 
enable airbag suppression for occupants below a 
specified seated weight. This technology could also 
be used to classify occupants into weight categories, 
thus improving the adaptation of restraint force for 
various impact situations. Transponders or ‘tags’ are 
also available for use with rearward facing child 



seats. They can thus be detected by a seat -based 
sensor and used to disable the airbag. Occupant 
sensing systems can work together with the SDM to 
control airbag deployment. 

Delphi’s Occupant Position and Recognition 
System (OPRS) recognises the passenger position as 
well as monitoring a specified zone around the airbag 
for occupant presence. The system determines whether 
a passenger is in the immediate proximity of the 
airbag at the time of a possible deployment and 
initiates appropriate action. Now under development 
it will initially detect empty seats, rearward facing 
infant seats and dangerously close occupants (pas- 
senger airbag only). Several sensing technologies are 
under consideration, including infrared, ultrasonic, 
thermal and capacitive. However it is likely that the 
final system will require a combination of these. 

System integration 

A complete system approach ensures that the limita- 
tions of current airbag systems will be remedied in 
the most efficient way. For example, the time at 
which the crash sensors have to make deployment 
decisions in order to properly protect the occupant is 
closely related to the characteristics of the airbag. 

This is also true for the desired classification of 
occupant weight and height based on the restraint 
system properties. Thus, sensors and airbags devel- 
oped together will utilise the full capacity of both, 
whereas a ‘system’ formed by individually devel- 
oped components may be less efficient and more 

Collecting, processing and utilising the occu- 
pant and crash situation data required for fully devel- 
oped adaptive restraint systems is a challenge to be 
met by developing a Safety Bus and Distributed 
Restraint System Architecture, based on the ‘plug- 
and-play’ concept used in the personal computer 
industry. The new system allows the addition or 
deletion of sensors and deployment loops (actuators) 
without redesigning the system. 

The Safety Bus concept uses two wires; each 
node derives its electrical power directly from the 
bus. The ‘arbitration’ bus architecture allows any 
device to initiate a transmission when the bus is in an 
idle state. The design accommodates two transmis- 
sion rates: high speed (500 kHz) for rapid system 
parameter updates and low speed (10 kHz) for diag- 
nostic/sensor information — seat occupancy, weight 
and seat belt status. Message throughput on the 

Fig 28: 



Crash Severity 






Seat Position 

Side Airbag 
Module & 

— Anticipatory 

i- Rear Facing 
Infant Seat 

i- Occupant 
and Posture 

- Frontal Airbag 
and Control 



company’s bus is eight times faster than existing 
CAN-bus protocols, it is claimed. 

A complete system approach also ensures that 
many of the limitations of the current airbags will be 
effectively remedied. For example, the time at which 
the crash sensors make deployment decisions in 
order to properly protect the occupant is closely 
related to the airbag’s characteristics. The desired 
classification of occupant weight and height based on 
the restraint system properties would work in a 
similar way. The company offers a complete line of 
highly integrated driver, passenger and side impact 
airbag systems and is an experienced manufacturer 
and integrator of steering wheels, dashboards, door 
modules, door trim and airbags leading to fully 
integrated modular cockpits. Products such as the 
Snap-In, Recessed and Invisa-ModTM Integral Steer- 
ing Wheel and Driver Airbag Module consolidate 
component function in this way. 

Attention to Leg Injuries 

A conventional airbag system supplements the safety 
belt in that it offers additional protection to vulner- 
able areas such as the head, the neck and the spine. 
However, lower body injuries have been put in focus 
recently because the introduction and use of safety 
belts and airbags have greatly reduced the severity of 
most upper body injuries and also because many leg 
injuries have been found to have long-term conse- 
quences. Traditionally, a ‘kneebolster’ is used: a 
padded structure that faces the occupant’s knees and 
applies restraint force to the occupant’s lower body 
through the femurs. In order to work properly, the 
knee-bolster must be positioned fairly close to the 

Fig 29: Pyrotechnic actuated venting 

occupant, but that also means it takes up valuable 
space and reduces the leg room for the driver and 

Hence, deployable knee-bolsters are now made. 
This makes it possible to move the dashboard for- 
ward in the vehicle, as the knee-bolster automatically 
positions itself in proper proximity to the occupants 
knees if a crash occurs. Recent development work 
has also shown that a properly designed, deployable 
knee-bolster has the potential to reduce lower leg 
injuries, keeping the legs in a more favourable posi- 
tion during a crash. 

The company offers an unique, low-mount 
airbag for application on the passenger side. It incor- 
porates a deployable knee-bolster, which is filled 
immediately when the bag deploys. An intricate 
chamber system then makes the gas fill different 
parts of the airbag at different times, thus eliminating 
the need for aggressive, immediate inflation of the 
whole system . The arrangement allows more free- 
dom in the design of the passenger area, not only 
because it eliminates the need for a knee bolster, but 
also because it wraps around and covers the instru- 
ment panel completely, helping to protect the passen- 
ger from potentially harmful surfaces. 

Controlling restraint deployment 

The ability to control airbag deployment force is a 
critical aspect of any adaptive restraint system. Del- 
phi’s Pyrotechnic Actuated Venting (PAV) airbag 
module utilises a slide mechanism to direct inflator 
energy into the airbag. Fig 29. When combined with 
developing sensor technology, PAV will provide full 
or ‘lower-power’ performance based on sensor input 
on occupant and crash characteristics. The system — 
available for both driver and passenger applications 
— combines the advantages of variable inflation 
with the cost-effectiveness of a single output inflator. 

Multiple inflators and pyrotechnically control- 
led venting help maximise occupant protection and 
minimise injury risk by adapting airbag inflation 
speed and stiffness to the occupant characteristics 
and crash situation. A new generation of algorithms 
for these technologies is being computer simulated 
and sled testing to evaluate the cost/benefit ratio. 
Airbag restraining stiffness can remain ‘HIGH’, as is 
the case with all current airbag systems, or can be 
modified to ‘NONE’ or ‘LOW’, based on various 



sensor inputs. Note that the ‘HIGH’ level airbag 
restraining force is only needed when the occupant is 
an unbelted adult and the crash severity is moderate 
to high. 

While fully adaptive restraint systems are not 
yet available, Delphi already produces a unique 
airbag design to protect out-of-position passenger- 
side occupants from injury during airbag deploy- 
ment. The patented Bias Deployment Flap airbag 
incorporates a box-shaped guide element, a ‘flap’ of 
material that diverts the deployment direction of the 
airbag away from an occupant located too close to the 
deployment surface. Through the Aegis joint ven- 
ture, they are able to offer the D-60 driver airbag with 
pyrotechnic, non-azide inflator. The use of non-azide 
propellant means that it creates virtually no residue 
or toxic gas. 

The Hybrid Inflator System, which was co- 
developed with OEA, provides a smoke-free gas to 
inflate the cushion; deployment is aided by a small 
pyrotechnic charge. Cushion inflation is precisely 
controlled for specified occupant energy manage- 
ment and a future adaptive inflation technology will 
provide dual-stage inflation of the airbag. 

Safety seating developments 

Other innovative restraint systems by the company 
include the Pro-tech self-aligning head restraint which 
has been specified for the latest Saab 9-5. It is 
designed to reduce whiplash injuries during low 
speed rear impacts, by lifting and moving forward 
and upward into the optimum position for occupant 
head and neck protection. The system is also avail- 

Fig 30: Catcher’s Mitt seating 

able as part of the so-called Catcher’s Mitt Seating 
from the company to be introduced this year, Fig 30, 
where it combines with a high-retention seat and the 
ABTS integrated seating and seat belt system in 
which the belt is attached to and moves with the seat. 
The seat absorbs energy during rear impact by allow- 
ing the pelvis and lower back to penetrate the seat 
squab in a controlled way, safely pocketing the 
occupant during the course of the impact. The Pro- 
tech head-restraint is activated by the energy of the 
impact, rearward motion of the occupant actuating a 
pressure plate and system of levers to move the 
restraint. ABTS involves the use of a high-strength 
seat structure and floor attachment to absorb belt 
anchorage loads. It particularly suits open car loca- 
tions where no B-pillar anchorage is available, also 
removable seat situations or those where a large 
amount of adjustment is provided. A low-mass ver- 
sion is available weighing 25 kg. Bosch has an- 
nounced that its Smart Airbag System is being imple- 
mented in three stages, the end of 1997 seeing start of 
production of the SMINC smart inflation concept to 
control the sequence of inflation according to the 
particular crash situation. Currently being introduced 
is the Automotive Occupancy Sensoring (AOS) sys- 
tem seen in Fig 31 which offers adult and child seat 
occupancy and out-of-position detection devices. In 
a latter phase a pre-crash sensor will be able to detect 
the relative crash speed before a collision, triggering 
individual protection mechanisms before impact. 

Fig 31: AOS system 

Chapter 3: 

Powertrain/chassis systems 

The challenge of natural gas engines increases as emissions 
legislation toughens for petrol and diesel engines. However, 
the further development of the small HSDI diesel, and the 
direct injection gasoline engine, is showing, thus far, that they 
are up to the challenge. The future could also lie in revised 
combustion cycles such as the constant-pressure one; there is 
considerable scope too in revised valve arrangements on 
otherwise conventional IC engines. In transmission design, 
clutch-to-clutch automatics vie with electronically controlled 
CVT, while in chassis systems, the mechanics of roll-over 
has come under the spotlight again, in vehicle handling 
analysis, as has the computer- simulation of suspension linkage 



Powertrains: the next stage? 

Careful comparative studies are revealing some strong 
emissions advantages for natural gas engines, but 
LNG may be preferred to CNG for CVs. For lighter 
duty vehicles, lean bum natural gas competes with 
advanced diesel and gasoline concepts but the ulti- 
mate solution may be IC-engines operating in a 
restrictive speed band in conjunction with continu- 
ously variable transmission. 

Important work on rating particulate emissions 
in terms of environmental damage has been carried 
out by MIRA’s Powertrain & Emissions Technology 
team. Using the test set-up seen in Fig 1, a feel for the 
relative merits of gasoline, diesel and CNG engines 
is obtained in this respect. The research team under 
Simon Greenwood have shown that particulate nu- 
clei start to grow as they leave the combustion 
chamber and pass down the exhaust tracts, as tem- 
peratures drop below 300C. Soot particles are found 
to agglomerate into chains with hydrocarbons and 
nitrates/sulphates condensing on to them. Some 90% 
of the particle size distribution is below one micron 
as measured by the Scanning Mobility Particle Sizer 

Greenwood explains that particle sizing tests 
used a 20 cm diameter particulate dilution tunnel, 
with an exhaust dilution of about 20:1. The exhaust 
sources were various motor vehicles on a 50 kW 
emissions chassis dynamometer. The SMPS probe 
was a steel tube facing the aerosol flow connected to 
the SMPS inlet. Vehicle and tunnel were precondi- 
tioned by driving at 50 km/h and constant speeds 
were chosen as being representative of the constant 
modal conditions within the European emissions 

Particfe Mobility Irvnl 

drive cycle. These speeds ranged from idling to 120 
km/h. Two diesel vehicles were tested for this work. 
The vehicles had a normally aspirated IDI engine and 
a medium sized turbocharged IDI diesel with cata- 
lyst, Fig 2. All three gasoline-fuelled vehicles pro- 
duce few particulates at idle and low power condi- 
tions. Particulates that were formed did not show any 
identifiable distribution. At higher power conditions 
there were large particulate populations. Fig 3. 

Of the fuels tested, ‘Vehicle Petrol 1’ produced 
the fewest particulates. A significant increase in 
particulate numbers started as low as 30 km/h and 1.0 
kW. The peak position developed at around 34 nm. 
‘Petrol 2’ produced distributions that were narrower 
and only developed above 80 km/h and 7. 1 kW. The 
peak first formed at 25 nm, corresponding to the 
Nucleation Mode, before abruptly shifting to 70 nm 
at high powers. This shift occurred as the particulates 
started to aggregate and received condensate from 
excess fuel. ‘Petrol 3’ showed similar behaviour but 
at a higher particulate concentration, the higher 
power peak diameter being 40 nm rather than 70 nm. 

The two CNG-fuelled vehicles produce few 
particulates at idle and low load, Fig 4. However 
unlike the gasoline vehicles/fuels described above, 
the particulates that are formed at low powers tend to 
form an identifiable and typically bi-modal nature. 
The first peak is at 60 nm, the second at 120 nm 
diameter. At higher powers a smaller diameter peak 
appeared and grew sharply to produce a very large but 
narrow particulate size distribution, with total 
particulate numbers similar to those seen in diesel 
and gasoline vehicles. In these tests, this vehicle peak 
occurred at 20-30 nm. 

The diesel vehicles produced similarly shaped 
distributions but with the normally aspirated engine 
producing rather more particulates. The difference 
may be the result of the catalyst on the turbocharged 
car. The diesels were the only vehicles to produce 
significant quantities of particulate matter at idle. 
The measured loaded peak positions were in the 
range 70-90 nm for the normally aspirated diesel and 
50 -60 nm for the turbo diesel. The particulates at idle 
were smaller at 32 and 46 nm respectively. The 
particulate diameters pass through a maximum in the 
medium power range before decreasing again. This is 
the result of the two competing processes. At idle, 
there are fewer particulates produced as there is less 
fuel delivered to the engine, meaning that the number 



of particulate cores is lower than under load. With around 50 km/h and 3 kW for two of the three 

fewer cores available, the particulate chains cannot vehicles. The growth in the distribution is rapid at 

grow as long as under load. In addition, the concen- road loads over 1 0 kW. In two of the gasoline 

tration of excess hydrocarbons is low so less conden- vehicles, this growth is paralleled with a shift to 

sation occurs. As the fuelling increases, the mean higher diameter, 40 nm and 70 nm respectively, 

chain length and condensation rate increase. How- clearly showing the bi-modal nature of the particulate 

ever, at high powers, although the number of size distributions as the shift is the change between 

particulates is very large, the exhaust throughput is the Nucleation mode and an Accumulation mode. It 

also rapid. This means that whilst agglomeration has been suggested that the engine oil also has a role 

does occur, it has less time to proceed to completion, in this behaviour, Greenwood points out. The 

In general the gasoline vehicles produced much particulate numbers at high power are similar to those 

lower particulate concentrations than the diesel at seen in the diesel experiments, 

low and medium powers. At higher loads a peak Generally, the natural gas vehicles produced 

forms in the range 20 to 30 nm, half that seen in the few particulates at low load. However there was a 
diesel experiments. The initial peak is formed at clear bi-modal distribution at this load. This indicates 

Fig 3: Gasoline particulate size 

Pamela Mobrtty Diameter (nml 



that there is a mechanism working to produce these 
particulates unlike in the gasoline vehicles. At high 
loads, the number of particulates increased very 
sharply. The high load distributions are very narrow 
and have peak values an order of magnitude higher 
than for diesel and gasoline although at a smaller 
diameter. This allows the total number of particulates 
under the distributions to be quite significant even in 
comparison to the diesel and gasoline vehicles. 

Greenwood concludes that all three fuel cat- 
egories produce significant numbers of particulates 
at high engine loading but at low loads it is the diesel 
which is the significantly heavier polluter and 
particulate size, too, is significantly higher. 

Truck and bus emissions 

In a seminar held by Millbrook Proving Ground, the 
results of tests on truck emissions were reported 
which were based on what the GM facility calls ‘real- 
world’ emissions testing. This is an important dis- 
tinction because current heavy-duty vehicle legisla- 
tion is based on engine-dynamometer rather than 
whole-vehicle testing. Using Millbrook’s VTEC fa- 
cility complete vehicles are evaluated for emissions 
under simulated operating conditions approaching 
field service. 

The Proving Ground’s Andrew Eastlake dem- 
onstrated the importance of using equipment such as 
VTEC since on choosing a catalyst formulation for a 
London bus, to give best results on the legislative 
dynamometer tests, this actually showed a conver- 
sion efficiency of only 40% under the ‘real-life’ 

Heavy Trucks 


Motor Cycles 

Light Trucks 

Fig 5: Road traffic NOx emissions, AD 2000 estimate 

1.0 1.2 1.4 1.6 1.0 1.2 1.4 1.6 1-0 1.2 1.4 1.6 


PwiiCtef Dimeter torn} 




Fig 4: CNG particulate size 



sient- operation emission data Eastlake found con- 
siderable cooling-off of the catalyst over long peri- 
ods, typified by long periods of idle experienced, say, 
in Oxford Street. On re-selecting a catalyst formula- 
tion to give best results on VTEC tests, 90-95% 
conversion levels were obtained. 

The crucial importance of truck and bus testing 
was also demonstrated by the projected emissions 
levels for different vehicle categories, shown at the 

Fig 7: BOC DS LNG-powered ERF 

Fig 8: Gas storage requirement for equivalent 
range , CNG (top) against LNG (bottom). 

Fig 9: Torque characteristics: diesel vs gas 

seminar by the pie chart of Fig 5, referring particu- 
larly to NOx emission in the year 2000. This is 
despite the far lower vehicle volumes in this category 
than those of passenger cars. In the case of buses, 
however, the engine emissions for passengers carried 
are much lower than those applying to passenger 
cars. With the European 1 3 Mode test for heavy-duty 
vehicle engines, testing is carried out at a series of 
steady speeds and loads, emissions being measured at 
each point. There is no measurement of transient 
acceleration effects nor cold-start emission effects. 
Because one type of engine is used in a variety of 
different heavy-duty vehicle types, the engine manu- 
facturer only needs to design and calibrate one engine 
iteration. Once this is homologated it can be sold to 
many different vehicle builders who are likely to 
optimize the installed engine for fuel economy and 
driveability rather than reduced emission levels. It is 
therefore proposed to adopt a new transient cycle test 
for Europe in AD 2000. However, it is still likely to 
be based on engine, rather than vehicle, testing. 

CNG’s advantage over gasoline 

Research at the University of British Columbia', 
carried out over a wide range of air/fuel ratios, has 
shown while generally power output is reduced by 
some 12% with CNG (compared with gasoline) due 
to the displacement of air by the gas, emission levels 
were substantially reduced. While both fuels exhib- 
ited nearly equal thermal efficiency, importantly 
natural gas showed increased efficiency at very lean 
mixture running, due to an extension of the lean limit 
of combustion over gasoline. The authors argue that 
by using a lean-burn strategy, with a carefully 
optimized SI CNG-engine, this might result in both 
light- and heavy-duty engines meeting current and 
proposed emissions regulations without the need for 
a catalytic converter. 

Tests were carried out in the university’ s Alter- 
native Fuels Laboratory in a Ricardo Hydra engine at 
three speeds and for one full-load plus two part-load 
conditions. Emissions results are shown in Fig 6 as a 
function of relative air/fuel ratio (RAFR) for gasoline 
(G) and natural gas (NG). At wide open throttle, HC 
emissions are some 50% lower with CNG while CO 
was 100% lower for all operating ranges. NOx levels 
have similar peak values, for both fuels, under full- 
load conditions but the advantage increases for CNG 
under part-load conditions. 



Liquified natural gas (LNG) 

The 12.2 litre Perkins Eagle LNG engine has demon- 
strated the effectiveness of natural gas in a real 
operating environment, used in ten ERF tractor units 
operating for BOC Distribution Services, Fig 7. The 
320 bhp engine is developed from the Perkins 2000 
Series of gas engines and Eagle TX range of diesels. 
The so-called TxSi unit is air-to-air charge cooled, 
fitted with a wastegated turbocharger and two-way 
oxidation catalyst. LNG has a fuel density over 
double that of CNG, allowing more energy to be 
stored in a smaller volume, Fig 8. LNG also has 
consistently high methane content, which allows 
reliable running in high compression ratio engines 
and gives more consistent emission control, say 

The IMPCO electronic air and fuel manage- 
ment system is a key feature of the engine. It simul- 
taneously measures the mass of both air and gas, in a 
direct manner and with very fast response, so that the 
unit returns very low emission levels even under 
transient operating conditions. It meets AD 2000 
legal requirements and, because LNG contains virtu- 
ally no sulphur, long catalyst life is predicted. Noise 
emission is also very considerably less and at idle the 
level is down by 7 dBA against diesel. As seen in the 
torque characteristic of Fig 9, the engine can lug 
comfortably down almost to low idle speed. 

Exploiting the low C0 2 
emission of small HSDI diesels 

Ricardo engineers 2 have shown the possibilities of 
using diesel propulsion for sub-B class cars which 
will emit less than 90 g/km C0 2 exhaust emission. 
Potential fuel consumption less than 3 litre/100 km is 
also mooted for a 1 .2 litre 4-cylinder engine with 70 
mm bore. Such a car would have weight of 750-800 
kg and power rating of 45-55 kW. Fuel injection and 
maximum cylinder pressures (140-150 bar) are both 
considerably higher than existing HSDI diesels and 
common rail fuel distribution, with 17 mm dia injec- 
tor, is suggested. 

According to the authors, the current industry 
‘standard’ 17 mm diameter fuel injectors are more 
suitable for truck engines than small passenger cars 
when packaging for diameter and length. Experience 
has proved that an injector mounted vertically and 
central in the bore between a 4-valve pattern typi- 
cally provides the best overall performance for HSDI 

engines. However, with a relatively large injector 
and small bore, the space available for valves was 
naturally at a premium. The critical section was 
between the fuel injector and inlet valve centres, Fig 
10. The maximum size of inlet valves was limited on 
one side by the clearance to the cylinder bore and on 
the other side by the requirement to achieve a coolant 
path between the inlet port and the injector, also 
between the valve springs and injector. The use of a 
separate injector sleeve, in preference to a cast boss, 
has allowed for the maximum possible inlet valve 
size whilst maintaining cooling and casting integrity. 
Together the inlet valves shown gave a total flow area 
of around 15% of the bore area which was shown by 
simulation to allow performance targets to be reached. 

Fig 1 1 shows three possible porting arrange- 
ments: (a) shows two directed inlet ports and a 
tandem exhaust port. This achieved a swirl ratio of 
2.5 Rs and was also unrestrictive, leading to high 
volumetric efficiency — but had the disadvantage 
that narrow water jacket sections under the ports 
would result in weak sand cores and potential casting 
problems. The small bore of 70 mm meant that it was 
not possible to separate, with coolant, the long inlet 
port wall from the exhaust port of the next cylinder, 



introducing the possibility of inlet charge heating, (b) 
shows two helical ports, one long and one short. This 
arrangement had similar performance but the added 
advantage that if inlet port deactivation was consid- 
ered, a better range and control of swirl was possible, 
(c) shows side by side helical inlet ports, found to be 
restrictive and also limited to a maximum swirl ratio 
of 2.1. Their structure does not improve cylinder 
head stiffness, unlike the tandem arrangements. 

The packaging of the fuel injectors considered 
was dependent on their type, Fig 12. Two types of 
common rail injector were investigated: internal and 
external solenoid designs. Both B and C had external 
solenoid designs where the solenoid is attached to the 
injector either in-line or at 90 deg to the body. With 
this type the solenoid envelope requires a space of 
approximately 30 X 70 mm; A, however, has an 

a) Directed ports b) Long and short helical ports 

Fig 12: Alternative injector installations 

integral solenoid which fits completely inside the 17 
mm body of the injector. The figure shows that an 
injector with an integral solenoid offered improved 
packaging into the 300 cc/cylinder engine. This 
injector allowed for more freedom of valve-train 
design and achieved the smallest overall package 
size in the cylinder head area. Another advantage was 
that, due to its relatively short length and top fuel 
feed, it was possible to design a novel fuel rail which 
would also function as an injector clamp on a multi- 
cylinder engine. 

The main constraints on the valve-train were 
that the four valves had to be close to vertical and the 
injector had to package in the centre of the bore. Fig 
13 shows the alternative valve-train layouts that were 
considered most feasible. DOHC direct-attack was 
the simplest and potentially least expensive valve- 
train, but required a 9 deg valve included angle to 
allow for the central injector. This meant that dead 
volume in the head increased, which dropped the ‘K 
factor’ below the minimum desirable level of 70%. A 
DOHC valve-train was feasible, however, when us- 
ing finger followers which allowed positioning of the 
camshafts away from the centre of the engine and 
provided adequate clearance to the fuel injector. 
Although this design would be more expensive and 
less stiff than the direct-attack design the vertical 
valve inclinations allowed the least possible dead 
volumes and hence the lowest predicted engine emis- 
sions. Both alternative SOHC valve-train layouts 
used bridge pieces to connect the inlet and exhaust 
valve pairs. 

The SOHC -rocker design had a low camshaft 
and hence potentially the smallest overall package 
height when also considering the camshaft pulley. 
However, this design would be more expensive than 
the SOHC-finger design because of the extra rocker 
shaft machining required. OHV-pushrod/rocker was 
considered feasible for a low cost design but the 
performance of the engine would probably be limited 
by the high inertia and low stiffness of the valve-train 
components, making it impossible to produce the 
accelerations required to achieve the desired valve 
performances. Electro-hydraulic operation of the 
valves was investigated briefly, and potentially the 
freedom to have variable control over valve timing, 
lift, and duration would naturally allow for better 
optimization of the gas exchange process including 
internal EGR. 



The main structure could be conventional, 
using separate cylinder block and head, or a 
monoblock, Fig 14. Aluminium was considered a 
good compromise between cost and performance, a 
possible engine structure being shown left , the alu- 
minium head and block being through bolted from 
underneath, which allows extra packaging space for 
valve-train and fuel equipment in the cylinder head. 
The pistons are shown running in the parent bore 
aluminium which would be surface treated to achieve 
acceptable wear resistance. A monoblock engine 
structure is shown right. The head and block are 
integral, and the structure is completed by a bolted 
bedplate. The more uniform loading and cooling of 
the upper cylinder would reduce bore distortion and 
hence reduce oil consumption and oil sourced emis- 
sions. Satisfactory manufacture of a monoblock 
presents some difficulties. Firstly, the cylinder bore 
cannot be through-honed due to the closure of the 
cylinder by the gas face, which means honing would 
stop short of the gas face. A piston with a deep top 
land would then be required, penalizing emissions 
due to the increased crevice volume. This honing 
problem could be solved by inserting a fully honed 
dry liner into the bore, the only penalty being a small 
increase in engine length and weight; also machining 
the valve seat insert bores into the gas face has to be 
done through the bore. However, using a dry liner 
means that extra clearance is available before the 

liner is inserted. With this arrangement it was possi- 
ble to achieve acceptable tool clearance with valve 
sizes large enough to achieve the required perform- 
ance. In order to achieve good fuel consumption, 
parasitic losses to engine ancillaries must be mini- 
mized. Mechanical drives to these components means 
they are driven at speeds in a fixed ratio to engine 
speed rather than at their most efficient operating 
conditions. An alternative to a conventional alterna- 
tor is to use a flywheel starter/generator which could 
be used to crank the engine at high speed for starting 
and then generate the required electrical output once 
the engine is running. High speed cranking would 
help to reduce start- up emissions and electrical drive 
input could improve idle stability and reduce emis- 
sions during critical phases of the drive cycle. This 
system would be more commercially feasible if the 
HSDI engine were used in a hybrid diesel and electric 
vehicle. With vehicle demands for greater electrical 
outputs the demands on alternators has increased. 
Alternator manufacturers are developing water cooled 
alternators which have higher outputs and are less 
noisy than their air cooled predecessors, the authors 
point out. 

A novel approach to packaging the ancillaries 
for a 1 .2 litre HSDI engine is shown in Figure 1 5. The 
air conditioning compressor is mounted at the front 
of the engine with its clutch being driven directly 
from the crankshaft. A belt from the crankshaft 

SOHC - Rocker 

Fig 13: Alternative 
valve-train layouts 

DOHC - Finger 



pulley would drive a water cooled alternator, also 
front-mounted. Electrical power steering would be 
used. This arrangement reduces the width of the 
engine and in a transverse installation would mean 
either increased frontal crash protection or the possi- 

sion configuration were considered, conventional 
transaxle and an integrated engine and transmission 
structure. A transverse mounted engine with front 
wheel drive was assumed in both cases. A conven- 
tional bolt-on transaxle transmission provides the 
opportunity to easily adopt a range of transmissions 
for the vehicle. These would probably include manual, 
automatic, CVT and AWD. It is also known current 
technology. An integrated engine and transmission is 
shown in Fig 16. This arrangement has the transmis- 
sion components housed in the cylinder block with 
the final drive on the split line between the block and 
the bedplate. This powertrain would be stiffer and 
hence quieter than a conventional design, and would 
also be lighter. Overall powertrain length is reduced, 
which allows the engine to be mounted more cen- 
trally and for the drive shafts to have equal lengths, 
reducing torque steer and part-count. The disadvan- 
tage of this arrangement is that it would be more 
difficult to have vehicles with a range of transmis- 
sions when compared to a conventional bolt-on 
transaxle transmission. 

Direct injection gasoline 

Ricardo engineers are also taking direct-injection 
gasoline engines closer to production realization 
with work on the optimization of the combustion 
system 3 . In-cylinder gas sampling has been used to 
obtain local air-fuel ratio measurements in the devel- 
opment of a suitable piston bowl design. The meas- 
urements are used to validate CFD codes valuable to 
the design process. Stable combustion at 50-60: 1 A/ 
F ratios has enabled unthrottled operation down to 



zero load. 

Relatively low cost electronically controlled 
common-rail fuel injection systems are spurring a 
renewed interest in stratified-charge combustion, say 
the authors, using the technique of early injection at 
full load and late injection at part-load. Reduced 
cold-start enrichment and improved fuelling control 
in transient operation are also in sight, with the 
eventual goal of achieving the theoretically possible 
fuel economy of a diesel coupled with specific power 
of a gasoline engine. Success will depend on efforts 
now being made to control NOx emissions in the lean 

The optimum-shaped piston bowl of the com- 
bustion chamber is one which can produce an ignit- 
able mixture at the spark-plug at a crank angle for 
phased combustion to occur, producing best trade-off 
between thermodynamic efficiency and HC/NOx 
emissions over a wide speed/load range. Ricardo 
have modelled the process in the four steps of Fig 17: 

(a) applying Bernoulli’s theorem to the fluid flow in 
the injector nozzle to calculate spray characteristics; 

(b) comparing spray penetration with piston position 
to predict time when impingement on the piston is 
reached; (c) fuel flow along the piston until it de- 
taches as a vapour cloud; (d) assistance by air motion 
in bringing the vapour to the plug in time for ignition; 
Fig 18 shows a dynamic visualization of the flow- 
field using CFD. 

Fig 19 shows an example result of the total 
calculation over the four steps, its use being in the 
quick screening of combustion chamber design op- 
tions. Here, four different start-of-ignition timings 
have been used in calculating the fuel spray trajecto- 
ries prior to piston impingement. After flow along the 
piston and departure as a spray-cloud, it is assumed 
the tumble motion, measured at BDC by CFD, is 
compressed into the bowl as the piston rises and an 
angular momentum calculation is used to estimate air 
velocity at the edge of the piston bowl and hence the 
transportation time for the cloud to reach the plug. In- 
cylinder sampling tests were carried out using flat- 
piston crown, square and spherical bowl shapes, 
start-of-injection being varied between 310 and 330 
deg and injection duration held at 20 deg, equivalent 
to a load of 2.5 bar BMEP. Resulting A/F ratio at the 
plug is seen in Fig 20. 










Fig 17: Fuel stratification and transport mechanism 

Fig 18: Dynamic visualization of in-cylinder flow field 

12 i . r t . 1 1 1 r- 

-40 -35 -30 -25 -20 -15 -10 -5 0 

Sample Valve Timing At Nteidmtm Lift (dag. ATDC) 

Fig 19: Fuel spray transport calculation result 



Fig 20: Effect of injection timing on A/F ratio at 
the plug 

Fig 21: Hydraulically-sprung connecting rod 




— VW\r- 

1 - - - 




— m 


JaX p 

. AX,, 

Fig 22: Ideal cycle for Deocp engine 

cycle: the future for diesels? 

Researchers at Newcastle University and the Dalian 
Maritime University in China 4 have hinted that the 
limitations on diesel development, imposed by se- 
vere exhaust pollution control regulations, may be 
solved by unlocking the cycle efficiency of the 
constant-volume cycle. 

A variable engine cylinder volume controlled 
by cylinder pressure has been achieved in a diesel 
engine modified by a hydraulic cylinder, Fig 21, 
inserted into the connecting rod. The diesel engine 
with oil-cushioned piston gives the unit its name of 
Deocp. It is characterized by high cycle efficiency, 
improved low-load performance and good starting 
ability. The theoretical combustion cycle for the 
engine is seen in Fig 22 (points 123451) overlaid on 
the conventional cycle (12 3451). 

The effective stiffness K (with corresponding 
constant pressure P t ) of the hydraulic cylinder can be 
altered to change the characteristics of the engine, the 
table of Fig 23 indicating that engine compression 
ratio £ and efficiency 7j increase as the engine load p, 
decreases. A solid connecting rod is given by/f = 2.26 
X 10 9 and here it is seen that efficiency does not 
change with load. 

In tests of a physical engine, account had to be 
taken of the fact that piston movement on the Deocp 
is not exactly dependent on crank angle, owing to the 
hydraulic spring, hence the measurement technique 
shown in Fig 24 was used by the authors. Tests were 
carried out at 100, 90, 75, 50 and 25% full load. 
Results in Fig 25 (the PV diagram) show the 9% 
increase in thermal efficiency at 25% full load and 
the 1.6% gain at full load. 

The authors attribute the gain in thermal effi- 
ciency to the piston of the Deocp reaching its highest 
position, at the end of the exhaust stroke, because of 
its inertial force. Improved scavenging results from 
small clearance volume, and high static compression 
ratio, especially at low loads. Swirl is also created at 
this point by the rapid increase in cylinder pressure. 

The authors have calculated that the combus- 
tion process finishes about 7 degrees of crank angle 
earlier than that of a comparable conventional diesel. 
In examination of the indicator diagram, in crank- 
angle form (Fig 26), it can be deduced that the 
cylinder pressure of the conventional engine (CE) is 



higher than the Deocp at the beginning of the com- 
pression stroke because the CE has a relatively large 
amount of remnant waste gas. After about 55 degrees 
of crank angle BTDC the Deocp ’s compression pres- 
sure increases faster because of its high compression 
ratio. During the expansion stroke, as cylinder pres- 
sure falls, the compression energy stored inside the 
hydraulic spring is released to the engine cylinder 

and the expansion line of the Deocp is resultingly 
higher. The diagram also shows significant compres- 
sion of the hydraulic spring occurs at about 4 degrees 
BTDC, where combustion starts, with maximum 
compression taking place 10 degrees ATDC. 

SFC results are shown in Fig 27 and the de- 
tailed test results in Fig 28. 

K (N/m) 

P k (bar) 

P\ (bar) 


/’ 2 



1 .49 x 1 0 7 















1 653 







2.26 X 10“ 



1 1 67 





11 67 


1 .895 



1 1.67 

1 .006 





1 .004 

2 548 


Piston amplifier 

Pushing rod 

ZZf DMC-101 


A A Displacement 


Fig 23: Ideal cycle efficiency 

Fig 24: Piston displacement measurement 

Crank shaft angle (°CA) 

Fig 26: Cylinder pressure vs crank-angle fj g 25: Cylinder pressure vs volume 

Diesel engine with an oil cushioned piston Comparison engine 

Load (%) 





















P\ (bar) 











n (r/min) 











A/i (kW) 











0, <%) 











8, (g/kW h) 











pi (bar) 











A(pi/pi ) 


46 29 



47 63 






N„ (kW) 











«,■ (%) 


33 64 

3 1 .42 







25. IX 

(g/kW h) 










334 7X 

r (°C) 







1 1.5 




t C l°C) 











Remark i„ = IX.47, K - 1 .68 X I0 7 N/m, A = 38 bar 

Fig 28: Full test-results 



Valve arrangements for 
enhanced engine efficiency 

Three recent contributions to the literature of engine 
design and development show how major changes in 
engine efficiency can be effected by different valve 
control systems and configurations 

Intake valve disablement 

According to Rover researchers 5 the technique of 
disabling airflow through one inlet valve (Rover 
Asymmetric Combustion Enhancement, RACE) is a 
way of generating stronger axial and barrel swirl to 
improve engine combustion efficiency. In conjunc- 
tion with EGR it improved efficiency from 3 to 7% 
and with lean bum from 7 to 11%. Because high 
levels of EGR or excess air normally lead to poor 
ignitability, slower combustion, instable combustion 
and unbumt HC emissions, a method has to be found 
of increasing the tolerance of charge dilution. Devel- 

Engine speed (rpm) 

opment work was therefore carried out on a 10.5:1 
compression ratio 4-valve per cylinder engine, fitted 
with air-assist fuel injection. 

Previous work had shown that increasing barrel 
swirl alone was not enough to achieve charge dilution 
tolerance and hence the decision to increase both 
axial and barrel swirl with an arrangement as seen in 
Fig 29 used for high activity, homogeneous charge 
control. Prior to disabling the second inlet valve, an 
inducer in the form of a port mask was fabricated to 
give barrel swirl ratio of 2.4 with both ports flowing. 
Flow testing of the cylinder head the gave the results 
of Fig 30. 

Air/fuel ratio loops at a 2000 rpm/2-bar BMEP 
test condition were obtained for the RACE engine, 
with and without the inducer, and compared with the 
base engine as shown in Fig 3 1 . A very good lean bum 
limit and large fuel economy benefit resulted, with 
the low-swirl engine working slightly better than the 
high swirl one. Generally, high activity of the RACE 
engine speeds up the early bum and stabilises lean 
combustion. Lean burn NOx limit emissions were 
also lower than the base engine although emissions 
were higher at stoichiometric due to the faster bum 
of the RACE system. Higher EGR tolerance and 
greater fuel consumption better than the base engine 
were also achieved. 

Hot-wire anemometry measurements were used 
to show that high levels of turbulence were present 
while maintaining low bulk velocities in the combus- 
tion chamber, believed by the authors to be the key to 
igniting a dilute mixture. 



Late intake valve closing 

Researchers at Sheffield Hallam University 6 have 
shown fuel savings as much as 7% can be obtained by 
the elimination of most of the pumping work energy 
loss inherent in late intake valve closing (LIVC) load 
control. They have also shown that, combined with a 
variable compression ratio device, LIVC can achieve 
up to 20% savings and a two-state LIVC control 
mechanism is under evaluation. 

Energy loss in pumping gas past the throttle, to 
the point where no useful output is generated at idle, 
is considered by the authors to be an important 
contributor to engine inefficiency. The LIVC con- 
cept achieves engine power modulation without the 
need for a throttle valve, trapped charge mass being 
reduced by keeping the valve open during a portion 
of the compression stroke, allowing excess induced 
charge to be returned to the intake manifold. Fig 32 
compares the throttled and LIVC engines by their PV 
diagrams, shaded areas depicting pumping losses. 
While conventional variable valve-timing mecha- 
nisms can alter valve phasing by some 30%, the 
authors consider that for pure LIVC operation altera- 
tions in excess of 100% are necessary. 

Earlier workers reportedly noted a slight loss in 
indicated thermal efficiency with LIVC, mostly 
caused by a reduction in effective compression ratio. 
Some workers used a secondary camshaft above the 
rockers, on a push-rod engine, to achieve LIVC of up 
to 90 degrees of crank angle. Other workers report- 
edly tackled the loss of thermal efficiency by adopt- 
ing the Otto-Atkinson combustion cycle, restoring 
the compression ratio at light loads by a variable 

102 - 

compression device for reducing clearance volume. 

The Sheffield Hallam approach of allowing a 
two-state LIVC control involves a relatively simple 
mechanism compared with that used by earlier work- 
ers, some of the load control being carried out with a 
conventional throttle. Seen in Fig 33, the arrange- 
ment consists of an intake camshaft with a profile to 
cause late closing of the intake valve at all times 
while for full-load operation a reed valve in the intake 
manifold prevents the charge being rejected from the 
cylinder. At low loads the reed valve is bypassed to 
allow the charge to return to the manifold. The 
authors point out that the reed valve must be posi- 
tioned as close as possible to the inlet valve while 
leaving sufficient space for the bypass. Future devel- 
opment is seen as incorporating the system within an 
Otto-Atkinson cycle engine 



Fig 32: Part load modulation by late intake valve closing 

Fig 31: Effect ofAFR on fuel 




Two valve Four valve 

Six valve Eight valve 


Single valve Three valve Five valve 


Multiple valve arrangements 

With the trend to increasing efficiency by the use of 
multi-valve combustion chambers, researchers at 
Queen Mary and Westfield College, London Univer- 
sity 7 , have developed a simple method of making 
preliminary assessments of such schemes, based on 
calculation of valve effective flow areas for cham- 
bers with eight or less valves. Seven valve heads are 
seen as an upper limit, with 5-7 valve systems 
meriting investigation in view of 4- valve heads being 
shown to have a 37% increase in valve area over the 
traditional 2-valve systems. 

Effective area is calculated from the steady 
flow equations applied to a constant pressure model 
for obtaining steady mass flowrate. The authors 
quote expressions for effective area in both choked 
and unchoked modes of flow. They obtain a non- 
dimensional effective area by dividing the above 
values by the generic area of the valve seat as 7cd 2 /4. 
The resulting values are compared in Fig 34/35 for 
different poppet valves as a function of valve lift. 

Maximum ideal area, generic area, of a single 
valve (Fig 34) is given by the authors as 

[R/(l + cosec a)] 2 n 

and for N valves the half angle a is n/N and the 
generic area ratio 

Fig 34: Generic areas for (a) single, and 

Fig 35:(b) even number and (c) odd ” ' 



where j = i or e, the ratio of valve and cylinder radii 
and area ratios being shown in the table of Fig 36; the 
generic flow area ratio of the different valve arrange- 
ments is seen in Fig 37. These two figures describe 
the authors’ key ideas, the last figure having curves 
intended only to show curves, just the points having 
physical meaning. 

The analysis above is based on inlet and ex- 
haust valves of identical head diameter and Fig 38 has 
been provided to show an extension of the analysis 
for unequal size heads. This three valve case has the 
same exhaust area as that for the two valve head while 
the inlet valves have a 58% combined larger area. 
The method could be used to optimize valve areas, 
possibly, the authors suggest, by slightly increasing 
exhaust while reducing inlet valve head sizes, to give 
significant overall gain. 

A special case for 5 -valve heads is made (1) 
with two inlet valves the same size as in the four- 
valve head, giving 37% increase in area over the two- 
valve case and (2) with three exhaust valves the same 
size as those for the six- valve case, giving a 33% area 
increase. These points are shown in Fig 39. 

Fig 38: Ideal 3-valve scheme 

Fig 39: Poppet valve 
effective area 



□ Inlet Valves 
A Exhaust valves 
# Inlet and exhaust valves, 

Total number 
or valves 

Half angle 
by one 
valve x 

Radius ratio 

Cases considered 

Area ratio 
for number 
oT inlet 

Area ratio 
Tor number 
of exhaust 

Number of 

n i 

Number of 



Practical value 





say 0.5 






















= 0.4142 





1 4- J7 














4 = 0.3333 









































4 = 0.3333 







Fig 36: Poppet valve half angles, radius ratios, area ratios and number of valves considered 



Trends in transmission design 

The design of the 4-speed all-clutch-to-clutch elec- 
tronically-controlled automatic transmission devel- 
oped by Honda for the 1998 Accord was described 
recently by company engineers 8 . Newly developed 
technology guarantees improved shift quality and 
fail-safe at high speed without dependency on one- 
way clutches as in the more conventional system it 
replaced. Already, by 1996, the company had devel- 
oped an all-clutch-to-clutch 3-shaft design of short 
overall length. Old and new gear train schematics are 
shown in Fig 40. 

In the conventional system seven engaging 
elements involve five multiple disk clutches, a 1-2 
one-way clutch and a servomechanism. The 
countershaft-mounted one-way clutch transmits driv- 
ing torque when accelerating in 1 st gear but it overruns 
without torque transmission when decelerating in 
this gear or when driving in other gears. In 1 st gear the 
holding clutch engages to apply engine braking for 
deceleration. The one-way clutch thus performs dual 
functions of improving up-shift quality from 1st to 
2nd and preventing 1st gear drive when a high speed 
malfunction occurs. 

In the new system there are five engaging 
elements comprising four clutches and a servomecha- 
nism. The 1st clutch transmits torque, while acceler- 
ating in 1st, to the 1st range and applies engine 
braking for deceleration. Cross-section of the new 
transmission is shown in Fig 4 1 and its specification 

3rd 4th 

1 st Gear Holding ( I 1 J I Reverse Main Shaft 

Torque Output to Tire 

Fig 40: Gear train schematics for conventional (left) 

in the table of Fig 42. The elimination of the two parts 
meant that two separate approaches to maintaining 
the performance criteria had to be taken. 

When the conventional transmission shifts up 
from 1 st to 2nd, engine torque output first is transmit- 
ted to the main shaft via the torque converter, then on 
to the idle gear, secondary shaft, 1st clutch, 1st gear, 
one-way clutch, countershaft and finally to the drive 
shaft. The computer starts to engage the 2nd clutch 
when 1st to 2nd gear requirement is indicated, and 
when 2nd-gear torque has increased to equal mainshaft 
torque, the transmission changes to the inertia phase 
and 2nd gear takes over all power transmission. At 
this point the one-way clutch overruns so that 1st gear 
does not contribute to power transmission and the 
shift quality is governed by the drive shaft torque as 
the product of mainshaft torque T m and ratios i of 2nd 
and final drive F gears. 

When shifting up from 1 st to 2nd gear with the 
new transmission, operation is the same as for the 
conventional one up to the point when the computer 
starts to engage the 2nd clutch. Then, when second 
gear torque has increased to the sum of mainshaft and 
lst-gear torques, the transmission changes to the 
inertia phase. As the transmission does not have a 
one-way clutch, the 1st gear contributes to power 
transmission and the drive shaft torque is given by: 

T m i 2 i F- T ,( i -^ i F 

The minimum drive shaft torque value, when 

T orqu© Output to Tire 

and new Honda ( right) automatic transmissions 



the clutch-to-clutch transmission changes from the 
torque phase to the inertia phase, is reduced by T J (i l 
- i 2 ) i r At this time the driver senses that the torque 
has dropped and the shift quality has deteriorated. 
Therefore, in order to prevent this deterioration in 
shift quality, it was necessary to establish a method 
for accurately controlling the 1st and 2nd clutch 
hydraulics at highly precise timing so that T ( = O 
when T m = T r Dealing with the occurrence of high 
speed malfunction in the conventional transmission 
involved a fuel cut operating as engine speeded 
above its set limit and overunning of the one-way 
clutch. With the new transmission a centrifugal hy- 
draulic cancellation device eliminated centrifugal 
pressure build-up by the rotating 1 st and 2nd clutches; 
the control system is reworked and two linear sole- 
noid added which switch clutch hydraulics at the 
appropriate timings. There is also a new shift-logic 
hydraulic circuit employing a fail-safe system, in- 
volving three non-linear switching solenoids, to pre- 
vent downshifting into 1st with an engine overspeed. 

CVT for 2-litre engined vehicles 

Fig 41: Honda automatic transmission 

Nissan engineers 9 have succeeded in commercializ- 
ing continuously variable transmission for the upper- 
middle class of car, and improving accelerative per- 
formance from rest over existing CVT systems by the 
incorporation of a torque converter which has also 
resulted in better creep capability. The electronic 

Fig 43: Nissan CVT for 2-litre cars 

Fig 42: Honda transmission specification 



control system has achieved improved fuel economy 
by attaining fine-tuned driveability and lock-up op- 
eration at low speeds; also a manual shift mode has 
been successfully incorporated. 

A cross-section through the new unit is seen in 
Fig 43 indicating the principal elements of (a) torque 
converter, (b) forward/reverse actuator, (c) ratio- 
change mechanism, (d) electro-hydraulic circuits 
and e. reduction gears. Specification and configura- 
tion are in Figs 44 and 45. The structural layout is 
similar to a conventional automatic transmission, 
only the change-speed mechanism being adapted to 
belt and pulley layout of a CVT. Overall length, 
including torque converter, is thus much the same as 
a conventional automatic gearbox for a 2-litre car. 

Less booming noise than that of a conventional 
automatic’s geartrain is inherent in the new unit, 
however, because inertia of the belt and pulleys is 
larger. The lock-up clutch has been given facing 
material with twice the durability of that on a conven- 
tional automatic to give lock-up speed of around 20 
kph compared with the normal 50 kph, providing the 
main fuel economy benefit. The ratio-change mecha- 
nism uses a newly developed 30 mm wide Van 
Doome belt, with nine layer steel band and 400 
friction elements. The additional width over the 
normal 24 mm gives the unit its higher torque capac- 
ity. Fig 46 shows the enhanced performance over a 
system used for a 1.3 litre Nissan car. 

The variable-ratio pulleys are arranged such 
that a tandem piston design on the primary pulley 
gives two-stage operation to generate the necessary 

force required in overdrive operation. On the second- 
ary pulley, because of its particular high speed opera- 
tion, a cancelling chamber is incorporated into the 
hydraulic cylinder; it prevents any loss in perform- 
ance due to centrifugal effects on the hydraulic fluid. 

High-torque manual transmission 

Engineers at New Venture Gear 10 described the de- 
velopment of the company’s NVT-750 gearbox, de- 

Forward/reverse actuation mechanism 

Fig 45: Nissan CVT configuration 

Applicable engine 

Displacement. 2.0L 

Max power: 190ps/7000rpm 

Ratio change range 

2.326 - 0 434 



Launch element 

Torque converter with 
low-speed lock-up clutch 

Forward/ re verse 
actuation mechanism 

Planetary gearset 
♦ wet multiplate clutch 

Ratio change 

30-mm-wide steel belt 
+ pulleys 

Oil pump type 

Hypotrochoid gear pump 

Final gear ratio 


Fluid type used 

Nissan Genuine ATF 
(special-purpose CVT fluid) 

_ " r 

Nissan belt-drive CVT 

1 .3— L class 


2.0-L class 

Belt working radius at Low ratio 
(Primary pulley) 

27.1 mm 


33.9 mm 

Pulley force at Low ratio 
(Primary pulley) 

28 9 kN 


37.6 kN 

Friction coefficient between 
elements and pulleys^ 



Pulley angle 



F x R 

783.2 kNmm 


1274 6 kNmm 

F x R ratio 




Fig 46: Torque capacity potential of belt-drive CVTs 

Fig 44: Nissan CVT specification 



Fig 47: New Venture Gear NVT-750 gearbox 

signed as a high torque density transaxle suitable for 
heavy duty FWD operation. The development objec- 
tives were the achievement of greater durability than 
the 650 unit while achieving world-class gear noise 
reduction targets. The 5-speed unit has torque capac- 
ity of 290 Nm, synchronization on all forward gears, 
13.08 maximum reduction and a 200 mm distance 
between input and output, Fig 47. 

Greater gear durability was obtained chiefly 
with a 15-20% increase in face- width but this was 
obtained without any change in the length of the 
gearbox over the previous model. This was achieved 
by developing net-forming clutch teeth tucked inside 
the gear width and reducing shaper cut tool clearance 
by going from one to two-piece welded construction, 
Fig 48. Gears were designed with total overlaps less 
than three times the sum of helical and profile; some 
10-15% compressive stress reduction resulted. 

The bearing installation was redesigned so as to 
increase effective shaft rigidity, the input shaft now 
being supported on a roller-ball combination de- 
signed to negate the effect of differential thermal 
expansion between aluminium alloy and steel with- 
out loss of support. The front intermediate shaft 
taper-roller bearing is now fitted with a high tempera- 
ture polymer cage which allows more rollers to be 
included within a given space envelope, despite the 
narrower width constraint placed by the increased 
gear face-width.The taper bearing cup is first pressed 
into the case, during assembly, the bearing then 
seated in the cup and, finally, the whole intermediate 
shaft assembly pressed into the taper bearing cone 
bore without brinelling the rollers. Easy gauging of 
the bearings is thus permitted with a simplified 
assembly process. 

The dual-cone synchonizers have organic fric- 
tion material, the 3/4th and 5th synchonizers being 
arranged with a common blocking ring, giving a 
claimed 400% increase in energy absorption over the 
conventional brass ring set-up. The reverse idler is 
engaged between the input shaft and an external spur 
gear on the 3/4 synchronizer sleeve, torque transmit- 
ted resulting in lateral motion of the sleeve during 
drive and coast reverse engagements. This motion is 
limited to 1-1.52 mm by using the 3/4 fork as a thrust 
surface, Fig 49; clutch teeth are located away from 
the sleeve by 1.27 mm to compensate for the remain- 
ing motion left in the sleeve, preventing premature 
engagement with the blocker ring. 

Fig 49: 3/4 shift system 



Fig 50: Base (above) and high level (below) strategies 

Control strategies for CVT 

A timely study of the control criteria for continuously 
variable transmission has been made by researchers 
at Ricardo-FFD' 1 who have provided a most lucid 
explanation of engine/transmission matching. They 
remind us that too little overlap in the torque curves 
for different ratios of a stepped transmission results in 
gaps in the shift sequence with engine speed, and 
hence power, having to drop too far between shifts — 
thus breaking the gear-change rhythm. Connecting 
the peak power points in a group of typical 2-D graphs 
of this sort would show a 100% efficient transmission 
characteristic working with an engine held at full- 
power continuously. When plotted in 3-D form, 
physical 3-D steps can be seen between the torque 
curves and these become a smooth surface with CVT. 
The surface has depth in the sense that part-throttle 
performance characteristics also exist and navigating 
through the resulting solid volume is the job of a CVT 
control system. 

The FFD PDSL simulator allows control sys- 
tems to be tried out on a sophisticated computer 
model of the powertrain. Control strategies analo- 
gous to those in-built into human responses for 
manual gearbox control can be tested, for example. 
Fig 50 compares PDSL interpretation of a base 
strategy ‘if in doubt move up the ratio range’ with a 
higher level one of ‘kick-down’ for possible overtak- 
ing, say. Similar strategies are also provided by the 
authors to demonstrate the operating requirements of 
racing cars and delivery vehicles, as extreme exam- 


Fig 51: Proposed CVT engine 
torque curve matching to wide 
range CVT 




Engine torque T |Nm| 



Accelerator p«d*l 


interface *d«. 


k rtf 


Control Unit 

Stored data. 

Engine performance gngl» 
Operating charattcnitic 


Cloir Open locK*up 
Clwtch charectcmcc 
( Blrvtaflen of 
vanjlv ratio 
eng i»c tniqac 
cociael pr**<unr 
dutch pretrorc 
clraung t opening of kick, 
op dutch 

Entioe jpeed 

Frruirt rrfibria( 

lock-ap cluck 

pulley (peed 

Ontpnl speed 

0*1 imp 
le wiper* fere 

Frm«rt ref wlaOwg 

• ratio 

• contact preuurc 

• ditch prmirc 

Fig 52: Vehicle system diagram 


Converter I 

CFT 20 E 
Eeotronlc | 






Car aanafaclmrer 

i, mm wl 



Driving I 

• ■» 







apcrmUsg point 





Fig 53: Drive strategy for influencing factors 

Eagioe speed n |rpo>) ► 

1 ■ F.coaomy range 
Power range 
K/y'-'yi Engine brake 

Fig 54: Adaptation ranges of drive strategy 

pies, the first wanting to hold maximum power 
through the accelerating sequence while the second 
is wanting a narrowing of the performance corridor to 
preclude cargo-damage. FFD believe a CVT strategy 
should be broadened from merely providing an infi- 
nite number of ratio selections to sophisticated en- 
gine/transmission interaction in a way that the driver 
sees no limiting characteristics of the transmission. 
The opportunity to map the control domain exists 
rather than moving from one rigid strategy to another, 
giving the driver hints and feedbacks as to the right 
way to use the transmission. Matching conventional 
engines to conventional transmissions requires, in 
the case of heavy haulage, for example, engine torque 
back-up for reducing the need to change gear. How- 
ever, this flexibility in power delivery is paid for 
because the engine is compromised between two 
duties, providing maximum power at maximum speed 
and maximum torque at usually a much slower speed. 
With a wide-range CVT there is no need for such 
flexibility and engine operating speed range can 
potentially be considerably narrowed down. This 
scenario has been modelled in PDSL, Fig 51. 

In the case of the German maker ZF’s CFT20 
CVT, flow of signals between driver, engine and 
transmission is seen in Fig 52. Fuzzy logic control is 
involved and the ECU is adaptable to different driver 
strategies. Achieving optimum drive characteristics 
was done by arranging for automatic specification of 
the engine operating point, accounting for the special 
features of each driveline. With the extensive range 
of possible system modes available an adaptive strat- 
egy is necessary and this led to the use of fuzzy logic 
in specifying the operating point. Fig 53 showing the 
system structure. The operating point is generally 
fixed by the need to optimize fuel economy and is 
only adapted to higher engine speeds (more power) 
by driver demand or operational task such as gradient 
negotiation. Converting performance demand into 
increase in operating point involves the evaluation of 
many parameters; for example, a distinction being 
made between setting off in an urban environment 
and overtaking out-of-town. Performance level evalu- 
ation takes the form of an adaptation factor infinitely 
variable between minimum fuel consumption and 
maximum road performance. A large adaptation area 
is also covered by the range of engine characteristics, 
plus accounting for all driveability requirements, 
Fig 54. 



The mechanics of roll-over 

As economy vehicles for urban operation become 
shorter, and therefore taller so that volume capacity 
and visibility are not compromised, the problem of 
roll-over intensifies. The appearance of new software 
from American Technical Publishers for allowing 
calculation in the fundamentals of vehicle dynamics 
gives an opportunity to better predict likely vehicle 
performance in this respect at the design stage. 

According to Gillespie 12 , who originated the 
software, the initial simplified quasi-static behaviour 
of the vehicle in cornering can be studied with the 
elementary model of Fig 55. The vehicle is assumed 
to be in a steady-state turn without any roll accelera- 
tion and the tyre forces shown represent the sum of 
front and rears. Super-elevated (banked) highways 
help to resist the overturning effect and the symbol j 
for the road-banking angle is used in calculation as 
positive when tilted towards the cornering vehicle. 
Lateral acceleration in gravities can thus be ex- 
pressed as: 

((1/2) + <prti- ( F/Mg)t}/h 

and the limiting value when load on the inner wheel 
becomes zero is: 

{(t/2) + fh}/h 

thus on level roads only the CG height and track are 
necessary for a first estimate of resistance to roll- 
over. Though this term gives conservative values for 
propensity to roll-over it is useful for comparison 
purposes and the author gives the table of Fig 56 to 
illustrate this. Since the peak tyre-to-ground friction 
co-efficient is quoted as 0.8, the table suggests that 
the lighter vehicles are safe from roll-over as they 
would slide out of the curve, which of course has been 
disproved in practice. 

In fact, by neglecting the compliances of the 
tyres and suspensions, roll-over threshold is overes- 
timated as the lateral offset of the CG is unaccounted 
for. Thus by considering the more refined model of 
Fig 57, incorporating the hypothetical roll-centre, 
lateral acceleration, related to roll rate R,, becomes: 

(t/2h){l/[l + RJ1 - H/h)]} 
the curly bracketed term showing the factoring roll- 

over threshold, usually around 0.95 the author ex- 
plains, but quite a lot higher for a low-slung sports 
car. Solid axles, with high roll-centres, reduce the 
effect of lateral CG shift compared with many inde- 
pendent suspension systems with roll-centres at near 
ground level. Lateral deflection of the tyres adds a 
further 5% reduction of the threshold, it is explained. 

However, for an accurate assessment complete 
modelling of the vehicle suspension is required ac- 
counting for such factors as lateral shift of roll-centre 
and lateral movement of the line-of-action of tyre 
vertical force, due to cornering forces and deflections 
arising from changes in the overturning moment 
under combined cornering and cambering effects. 
Such modelling would result in quasi-static roll 
response, indicated by Fig 58. This shows the linear- 
ity of the response breaks when one of the inner 

Fig 56: Roll reactions on suspended vehicle 

Lateral Acceleration (g) 



Vehicle Type 

CG Height 


Rollover Threshold 

Sports car 

18-20 inches 

50-60 inches 

12-1.7 g 

Compact car 



1. 1-1.5 

Luxury car 



1.2- 1.6 

Pickup truck 



0.9- 1.1 

Passenger van 



0.8-1. 1 

Medium truck 




Heavy truck 




Fig 57: Threshold comparisons 

Roll Angle, $ 

Fig 58: Equilibrium lateral acceleration 


Roll Angle 

Dt . 

Lateral Acceleration Input 

■ ■■ J 



Fig 59: Roll response to step input 

Fig 60: Threshold as a function of sinusoidal steer 

wheels lifts off. The plot suggests that highest roll- 
over threshold will be achieved by maintaining sprung 
mass roll rate at the highest level by using high roll 
stiffness suspension and designing so that simultane- 
ous lift-off of front and rear inner wheels occurs at the 
cornering limit. 

Transient roll-over 

To examine vehicle response to rapid changes in 
lateral acceleration a transient model is required. The 
simplest one would be that of Fig 57 but with a roll 
moment-of-inertia term added. This is useful for 
responses to step inputs such as changes in road 
surface friction or release of brakes after wheel-lock, 
the response of the system being similar to a single 
degree of freedom system, Fig 59. The fact that roll- 
angle can overshoot, depending on the degree of roll 
damping, the author explains, means that wheel lift- 
off may occur at lower levels of lateral acceleration 
input in transient manoeuvres, hence a lower roll- 
over threshold. 

In the case of sinusoidal lateral acceleration 
input, as approximated by the slalom ‘moose’ test, 
roll-over threshold response is shown in Fig 60 as a 
function of frequency for different classes of vehicle, 
the values approaching the steady-state cornering 
case at zero frequency. Heavy trucks are seen as being 
particularly sensitive to this effect since their roll 
resonant frequency is often less than 1 Hz and lane- 
change or certain roundabout manoeuvres of around 
0.5 Hz could result in roll-over. Cars having lower 
CG positions usually have roll resonances at about 
1.5 Hz. 

The most complete and accurate picture is 
provided by models which simulate both yaw and roll 
response. Using such a model to examine sinusoidal 
steer inputs reveals the important additional effect of 
front and rear tyre force phasing lag caused by the rear 
tyre not building up side force until a slip angle builds 
up. Fig 61 illustrates the effect. Here the front to rear 
lateral force build up lag is 0.2 seconds for the 1 Hz 
steer input case. The lateral acceleration for the 
whole vehicle is diminished from 0.8 to 0.5 g and to 
the driver there is a perception of lack of responsive- 
ness in transient manoeuvres. Four wheel steer cars 
eliminate this lag and improve responsiveness; how- 
ever, the author points out, it could contribute to roll- 
over propensity. 

In tractor-trailer vehicle combinations the phase 



lag is particularly pronounced and the response in Fig 
62 is typical. Here a 2 second sinusoidal steer input 
excites both a rearward amplification of the yaw 
response and roll-resonance of the full trailer, such 
that it experiences much larger lateral acceleration 
than the tractor, creating a dangerous whip-lash 
effect. The effect is prevented by using a tractor- 
trailer coupling which transfers roll-couples from 
tractor to trailer. 

Software example 

Gillespie used the quasi-static model to demonstrate 
the effect of highway super-elevation on vehicle 
occupants. The objective here is to find what angle 
would make the occupants experience 0.1 g lateral 
acceleration in a 40 mph steady-state turn of 500 ft 
radius in a vehicle of 60 inches track, 20 inches CG 
height and weighing 2200 pounds. 

Fig 63 shows the data input spread sheet with 
these values entered. On solving, the sheet generates 
the first three views of Fig 64, the third of which is a 
vector diagram showing the relationship between the 
vectors ac and g (relative to horizontal and vertical) 
and ay and az (relative to the vertical axis). The point 
common to the upper left comers of both rectangles 
represents the CG location while the line from the CG 
to G is centrifugal acceleration ay and the vertical 
line from CG to the other G is gravity acceleration g. 
On the other rectangle the line slanting right and 
upwards from CG to point V is lateral acceleration 
experienced by the occupants ay and the one slanting 
right downwards to the other V is the vertical accel- 
eration they experience. 

Neutral speed is that at which the occupants 
experience no lateral acceleration relative to the 
vehicle and the program allows back-solving for this 
case by changing (p from an output to an input 
variable, changing the input instruction for ay and 
blanking the V input field. This produces the sheet of 
Fig 65 and the fourth view of Fig 64, a vector diagram 
in which G = global axes (ac andg); V = vehicle axes 
(ay and az). 

The neutral speed for this curve is 29 mph and 
the equilibrium lateral acceleration plots are the 
same for this solution because the radius, super- 
elevation, track and CG height are unchanged. How- 
ever, the vector plot is different because centrifugal 
acceleration ac is smaller and the lateral acceleration 
ay is now zero. 







Fundamentals of Vehicle 

by Thomas D. Gillespie 

Chapter 9. Rollover 

Quasi-Static Rollover of Rigid 

Eq. 9-1, p. 311, and Fig. 9.2, p. 








Radius of turn 




Lateral acceleration (horizontal) 




T rack width (tread) 




Height ol CG 




Weiaht of vehicle - M’a 




Acceleration of gravity {default: 1 




Mass of vehicle 




Acceleration aionq vehicle y-axis 




Acceleration along vehicle z-axis 




Normal lore© on inside lire 




Normal force on outside tire 




Road cross-slope angle 

Rollover thresholds: 








Lateral acceleration (horizontal) 

ay roll 



Acceleration aionq vehicle y-axis 




Acceleration along vehicle z-axis 

Fig 63: First input spreadsheet 







Fundamentals of Vehicle 

bv Thomas D, Gillespie 

Chapter 9: Rollover 

Quasl-Stalic Rollover of Rigid 

Eq. 9-1, p. 311, and Fig. 9.2, 
p. 313 








Radius of tum 




Lateral acceleration 




Track width (tread) 




Heiqht of CG 




Weight of vehicle * M*q 





Acceleration of gravity 
(default: 1 g) 




Mass of vehicle 




Acceleration along vehicle y- 





Acceleration along vehicle z- 




Normal force on inside tire 




Normal force on outside tire 




Road cross-slope angle 

Rolover thresholds: 








Lateral acceleration 




Acceleration along vehicle 




Acceleration along vehicle 

Fig 65: Second input spreadsheets 

Lateral Tire Force (N / 1000) 



1.4 i 

1.2 j 


S' .8-: 

6 -) 


. 2 1 



Equilibrium lateral acceleration (cross-slope = 0) 

1 “1 I i 1 1 “ J ‘l 

5 10 15 20 25 30 35 

Roll Angle deg 

Equilibrium lateral accelerations ac and ay on cross-slope phi 

Fig 64 : Graphical output 



Suspension and steering 
linkage analysis 

For light vehicles, advances in modelling techniques 
are making the analysis of handling behaviour a 
much more realistic process than was possible with 
classical quasi-static techniques. The same is true of 
analysing suspension behaviour with respect to road 
damage by heavy vehicles. 

The change from active suspension systems in 
Formula One race-cars brought about by the 1994 
revised regulations focused a new interest in the 
perfection of passive systems. According to Brescia 
University researchers 13 , double-wishbone configu- 
ration, with either tie-bar or compression-strut brake 
force reaction, remains the favourite means of con- 
trolling wheel-to-chassis relative movements impor- 
tant for maintaining roll-centre location and stability 
of the tyre-ground contact patch. 

Continuing conflict existed, the authors main- 
tained, between low roll-centre to reduce lateral 

scrub at full-bump spring deflection, as well as large 
weight transfer and roll angles in cornering, against, 
with high roll-centre, reduced weight transfer but 
increased lateral scrub and the jacking effect. Key 
objective in design was to restrict the movement of 
the roll axis to a minimum to ensure constant levels 
of weight transfer in different operating conditions. 
FI practice is to adopt long parallel wishbone arms in 
order to minimize camber sensitivity to change in 
suspension ride height. 

Nome filedati: 

Denominazione : 




Altezza vettura: 


F94 wide, high RC 
-3.000 gradi 
000 gradi 
105.646 mm 
685.003 mm 

Premere un tasto per continuare 
Fig 67: Orthogonal projection of Fig 66 



The authors have developed MMGB design 
software for race-car suspension, in conjunction with 
Dallara Snc, for a complete 3-D analysis. Fig 66 
shows a prototype configuration involving a 16-node 
kinematic system, the wheel, hub-carrier and spring 
rocker lever being assumed to be infinitely stiff. The 
wishbones are each represented by two pin-ended 
members connected at the outer node. The spring- 
damper unit is modelled as a variable length beam 
while the spring-reaction and steering arms are con- 
sidered to be rigid. The program permits the linkage 
to be seen in orthogonal projection as seen in Fig 67. 
Data input files handle spring rate and road loads and 
the program is based on algorithms for kinematic 
solution and force equilibrium. 

The kinematic module computes geometry vari- 

camber (deg) 

ation during wheel travel and full bump rebound 
situations can be examined with the chassis effec- 
tively held still. Main suspension property outputs 
are seen in the graphs of Fig 68. A graphic animation 
is also available to simulate the motion of the suspen- 
sion for observation by the designer. The force mod- 
ule computes link loads, rocker-pivot loads and spring 
force for given wheel displacement and external 
loading. There is also a compliance module which 
effectively measures the interaction between the first 
two modules but taking into account link elasticity. 

Additional modules have been developed since 
the program was first launched. These allow study 
ofwheel rate and sprung mass frequency variation as 
the vehicle progresses along its track, arising from 
rising-rate or preload effects, also the effect of reduc- 

60 | 


■=■ -X 40 

E '■ 

s 20 k 

-40 1 

roll centre height (mm) 

toe in (deg) 

Fig 68: Main suspension geometry outputs from analysis 

spring length (mm) 



ing fuel weight and aerodynamic downforce inputs. 
Anti-dive and anti-squat configurations can also be 
examined. A roll-centre module can examine not just 
vertical but also lateral motions of the roll-axis. The 
very considerable lateral movement for a nominally 
ground-level roll-centre suspension design is seen in 
Fig 69 for a chassis roll of 4 degrees. 

The authors also referred to the work on sus- 
pension compliance of John Ellis whose book on 
vehicle handling dynamics was published in the same 
year 14 . Ellis recommends a vectorial approach to 
suspension analysis which is not dependent on loca- 
tion of roll-centre and which can be applied at any 
vehicle attitude. For two-dimensional analysis, in 
Fig 70, the origin O is fixed in the vehicle body plane 
and an outboard point O is at the wheel-ground 
contact. His analysis considers the velocities of these 
points in relation to the body and wheel velocities, 
relating them by the rigid link which effectively 
exists between the inboard and outboard points. 

Referring to the co-ordinates of these points, 
shown in the figure, velocity of the inboard point is 

V a = (~pz)j + (W + py)k 

That of the outboard point is composed of the wheel 
velocity V' and a component p due to rate of camber 
rotation of the wheel in 

V d = (v-p'z)j + pk 

the two velocities acting at the opposite ends of link 
AD. Since the relative velocities of these points along 
the line of AD is zero, the scalar products of these 
velocities with the unit vector of the link must equal 
each other, Ellis explains. By representing the link as 
unit vector 

ad = mj + nk 

where m and n define the spatial position of the link 
defined by co-ordinate geometry. Velocities of body 
and wheel/road interface then become related by the 

(-pz)m + (W + pyjn = (V - p z d )m-p’y d n 

By repeating this procedure for each link con- 

tw Velocity, outboard point. 

Velocity, Inboard point. 


Fig 70: Two-dimensional analysis 



Fig 71: Compliance representation 



necting body and wheel, a set of equations is obtained 
which can be solved simultaneously by matrix meth- 
ods to give useful design relationships such as the rate 
of contact patch lateral velocity with bounce velocity 
of the suspension; tyre scrub due to roll; and camber 
change due to bounce and roll. 

Ellis also carries out a three-dimensional sus- 
pension analysis using these techniques and finds the 
interrelationship with the steering system, as well as 
modelling the effects of tyre stiffness. Compliance 
effects are examined for the flexible mountings of the 
suspension. He defines the change in suspension 
characteristics due to external forces as the compli- 
ance and describes its quantification using the two- 
dimensional representation of Fig 71. 

Forces on the inboard bearings are found by 
static equilibrium equations, the stub axle O CD 
being under the action of normal and lateral forces at 
the road/tyre interface, which can be expressed alge- 
braically for this purpose. The line of action of the 
force within the suspension link AD intersects the 
force vector at a. The vector 0 3 is known, so also are 

the directions of the force vectors for joints D and C. 
The lower suspension link is under the action of 
spring force along FE and the forces at B and C. 
Directions of the reaction C and the spring force are 
known and intersect at (3, so the direction of the 
reaction at B can be obtained and the vector diagram 
drawn. By inputting the stiffness of the joints 1 and 
4, the deflections of these joints can be obtained and 
new positions of the inboard nodes of the suspension 
linkage obtained and the change in handling behav- 
iour calculated. A similar analysis for a 3D suspen- 
sion model is given in the book. 

He uses the steady state roll angle of the 
suspension, derived from energy and work equations, 
and expressed in terms of the suspension derivatives, 
to examine roll attitudes and load transfer. The roll of 
the body causes potential energy to be stored in the 
springs and work is done by the tyres in developing 
lateral forces to keep the vehicle in a comer; and 
developing lateral motion as a result of the roll. As 
the vehicle moves on its curved path, the lateral 
inertial force acting at the centre of mass gives rise to 
an overturning moment, transferring normal load 
from inner to outer wheels. Because redistribution of 
normal forces causes the lateral forces generated by 
the tyres to change in a non-linear way, the balance 
of the vehicle may change, Ellis explains. 

Roll-induced steering is possible at both front 
and rear axles, the relative roll stiffness of the suspen- 
sions affecting the proportion of the load transfer 
carried by each axle. Added to this are the effects of 
steering of the road wheels under roll-displacement, 
components of the lateral velocity at each contact 
patch due to tyre scrub also the camber of the wheels 
with roll displacement. Fig 72 (top) shows a 2- 
dimensional suspension for the case in which roll is 
the only motion and it is seen that the kinematic 
design of the suspension will constrain the motion of 
the road wheels in response to the roll of the vehicle. 
The kinetic energy of the system due to roll is 

2T = Ip 2 + (mV 2 + Ip' 2 ) 

but as V = dy/d<l>.p and p = df/d<p.p, suspension 
derivatives referring to the centres of mass of the 
suspensions, then for the whole vehicle. Fig 72b, the 
effective inertia in roll is 

/ = / + Em(dy/d<t>) 2 + EI(d<^/dp) 2 



Current review of road-friendliness ‘Road-friendliness’ assessment tests should: 

Cebon joined with colleagues from Warwick and 1 
Nottingham Universities to give an up-to-date as- 
sessment of road-friendliness research 15 with particu- 
lar reference to the common case of the articulated 
HGV. Recent thinking is that the dynamic compo- 
nent of tyre forces is the main factor in contributing 
to road damage. Not only is this significant compared 4 
with mean (static) tyre force but it is believed that 5 
different vehicles generate similar spatial patterns of 6 
dynamic loading. At a given speed, the tyre force 
time histories generated by a particular vehicle are 8 
repeated closely on successive runs over a given g 
stretch of road, the phenomenon of spatial 10 
repeatability. This means that certain points along 
the road will be subjected to the peak dynamic tyre 
forces generated by many vehicles, thus significantly 
increasing the damage incurred at those points, the 
authors explain. Assessment procedures should there- 
fore follow the criteria shown in the table of Fig 73. 

As well as the dynamic loading effect on road 
damage discussed above, a road stress factor has also 
been used based on the so-called fourth power law: 

E[P9t)] 4 = (1 + 6s 2 + 3s 4 

where P(t) is tyre force; E[ ] an expectation operator; 

P the static tyre force; and ,v the co-efficient of 
variation of dynamic tyre force (the die). However, it 
must be noted that the stress factor does not account 
for the spatial distribution of tyre forces and is no 
more than a ‘plausible rule-of-thumb’. 

The need to have road damage criteria that 
relate to specific points along the road has involved 
summing the measured dynamic forces of each axle 
raised to a power n, applied to each location along a 
road, or sensor in a road measuring mat. This is 
known as a weighted aggregate force model, the nth 
power aggregate force A" measured at location k 
being given by 

Z H (P;) for k = 1,2,3,..., N 

the power of n being chosen to represent the type of 
road damage being considered (4 for fatigue damage 
and 1 for rutting). The die and aggregate force 
methods have been compared in tests on a trunk road, 
Fig 74 showing the die averaged over all suspensions 
in each class of axle group for all trailer axle groups 

account for the effects of tyre and axle arrangements and static load 
sharing as well as dynamic loading performance; 
use testing equipment which is not dependent on the arrangement of 
axles and tyres, and therefore does not bias vehicle design towards 
particular configurations; 

3 be based on assessment criteria which reflect the potential of the 
vehicle to inflict fatigue and rutting damage to representative 
pavement structures; 

be conducted for representative conditions of speed road roughness 
and road construction; 
account correctly for wheelbase filtering; 

account for the effects of all axles of the vehicle on the pavement; 
resolve/sample correctly both high- and low-frequency dynamic 
tyre forces; 

utilize statistically significant samples of tyre force-time histories; 
account for the effects of spatial repeatability of dynamic pavement 

account for the dynamic interaction between the suspensions of 
tractors and trailers. 

Fig 73: Assessment procedure criteria 

Fig 74: Die results (2a and 3a refer to 2/3 -axle 
air suspension and 2a/3a to 2/3-axle steel 
suspension respectively) 

Fig 75: Fourth power results, error bars showing 
standard deviation 




for which suspension type was identified. Fig 75 
shows the fourth power results for the same trailer 
axle groups, normalized using the static weight of the 
axle group to enable different vehicles to be com- 
pared directly. The two criteria rank the suspension 
classes in different orders; this is due to spatial 
correlation between the tyre forces in an axle group 
which is not handled properly by the die. 

Generally it has been observed that dynamic 
wheel loads increase with increasing road speed and 
road roughness; that walking beam and pivoted spring 
tandem bogie suspensions generate the highest loads, 
leaf spring suspensions generate less, and air and 
torsion bar suspensions generate the smallest levels 
of dynamic loads. An important feature of dynamic 
loading is that all suspensions and tyres on a vehicle 
are involved in the sprung mass modes of vibration 
and so effect of tractor-trailer interaction is impor- 
tant. A trailer performing well with one tractor may 
perform badly with another. A practical assessment 
test is therefore likely to require a standard tractor for 
testing with every trailer and a standard trailer for 
testing with all tractors. 

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Chapter 4: 

Electrical and electronic systems 

The maturity of automotive electronics is seen in the desire to 
see the controller as seamlessly integrated with the vehicle 
system. This is evident in new developments in engine knock 
sensing, suspension, steering and drivetrain control. Vehicle 
system development is also being driven by telematic traffic 
management systems and the goal of vehicle headway control 
is not far away. Engineers from different disciplines are seek- 
ing knowledge of the digital systems used in electronic cir- 
cuits and applying the knowledge to driver- vehicle man-ma- 
chine control systems for creating information links with the 
surrounding road infrastructure and the driver becoming a 
predictable control link within refined electronic steering sys- 
tems. The greater sophistication in hybrid, electric and IC en- 
gine, drive for vehicles is seen in the more accurate matching 
of performance to the actual drive cycles which such vehicles 
are likely to experience. The use of map-controlled drive man- 
agement is emerging and some of the latest control techniques 
are seen in the first production hybrid, the Toyota Prius. Fu- 
ture possible hybrid development is considered by some to be 
with the use of supercapacitors to work in conjunction with 
energy storage systems based on batteries. Electronic control 
is even seen to be advancing into vehicle test engineering with 
the appearance of automated handling systems for track use. 



Automotive electronics maturity 

The seamless electro-mechanical vehicle is the cur- 
rent aim of automotive electronics engineers who 
seek to bring an end to inefficiencies caused by 
grafting on electric controls to systems originally 
designed for mechanical operation. This approach is 
seen here applied to engine, suspension, steering and 
drivetrain control. Meanwhile the principal growth 
area, telematics, is beginning to see firm proposals 
for OE navigation systems for less expensive vehi- 
cles in the adaptive cruise control and headway 
sensor described here alongside a proposal for a 
guidance system to allow vehicles to follow roadway 
white lines. Major proprietary system releases reveal 
the latest in engine management and rival systems for 
business driver communication with his/her office. 

According to Chrysler Advanced Technolo- 
gies specialist Dr Chris Bodoni-Bird 1 , the emerging 
technology of mechatronics is set to bring about the 
seamless electro-mechanical vehicle which can learn 
from Nature in its development. The theme is reduc- 
tion in mass by using seamless electro-mechanical 
integration by either exploiting intelligence and/or 
developing multi-functional componentry. This will 
also suit the increasing demand for new features 
within a fixed space envelope. Biometics is the study 
that sees how man-made systems can emulate Na- 
ture. In the key field of sensors and actuators perhaps 
man can learn from the sensing capabilities of certain 
bird species, for magnetism and light polarization, 
and electric eels and rattlesnakes for electricity and 
heat, he suggests. Exhaustive historical analysis of 
patents has shown that systems tend to evolve from 
mechanical through chemical and electro-mechani- 
cal to electro-magnetic (optical), and vehicle sensor 
development is said to reflect this. Because elec- 
tronic systems also have the capability for integra- 
tion, sensors can be shared between vehicle func- 
tions, as seen in Fig 1. Because, however, mechanical 
systems are often cheaper and difficult to replace, 
institutionally and technically, the compromise dis- 
cipline of mechatronics is seen by the author as set for 

A way of providing an on-board control system 
without the addition of hardware is described by 
researchers at NGK Spark Plug 2 who show how ion 
formation around the spark-plug can be used as an 
engine control technique. While the ion density 
represents the condition status of combustion, analy- 

sis of the condition is made possible by measuring 
voltage/current between plug-gap and ignition due to 
the magnitude of ion formation. In so doing the need 
is removed for sensors of EGR or knock and a plug- 
gasket type pressure sensor. 

While it is expensive to provide an ion detector 
for each cylinder, here the authors propose a tech- 
nique in which the spark plug electrodes are used for 
detection. But to avoid damage being caused by high 
tension voltages a point on the ion response curve, 
Fig 2, must be used after high-tension discharge. 
However, some increase in circuit complexity is 
involved, Fig 3. To measure ion density a coil is used 
to generate several 100 V pulses consecutively with 
the spark ignitions, the electrical circuitry providing 

Fig 1: Systems integration effect on sensors 

Fig 2: Ion current response 



the means to measure voltage decay with time, which 
is proportional to ion density, Fig 4. The ECU analy- 
ses combustion by examining the pressure-wave 
pattern (magnitude and pulse-width) using ion-den- 
sity and decay-period data. Input ignition signals, 
firing pulses, misfiring signals and the decay curve 
(and periods) are seen in the lower part of Fig 3. 
Amount of fuel injected is controlled so as to reduce 
the difference between fluctuation rate of the decay 
period and its standard fluctuation rate. Knock can 
also be detected because ignition and ion density 
waves are synchronized during detonation. But the 
decay period measurement technique alone cannot 
be used. The ion density voltage pulsation must be 
measured continuously. A capacitor is used to accu- 
mulate current generated by ion density voltage 
pulsation, and current flowing through an associated 
resistor in series with it, in order to sense the onset of 

Line connection 

Wave sequence 

-mg _n n n tl 

■3's»wi — n — 

Gnu, j"V 

n n a 

-sim, y i i it - ir 

Fig 3: Secondary voltage ionised current method 

Smart materials 
for suspension control 

Conventional mechanical suspension systems are 
potentially convertible to active control with the use 
of electronically monitored smart materials, in the 
opinion of researchers 3 at TRW Inc. Effective control 
of the motions of sprung and unsprung masses of a 
spring-suspended vehicle can only be achieved by 
interconnecting hardware responding and adapting 
to changing disturbances. Optimum performance is 
obtained, it is argued, when system energy can be 
decreased and increased as necessary. Achieving 
variable spring rate with conventional active suspen- 
sion systems requires high amounts of energy, with 
serious drain on the vehicle’s capability. 

By using smart materials, controllable by an 
external energy field, performance can be made to 
change in a predictable and repeatable manner. Among 
the smart solids, piezo-electrics exhibit a control- 
lable energy transformation between electrical and 

Fig 4: Determining flame resistance around plug gap 



mechanical modes, hence their use in noise and 
vibration control, the authors explain. Smart liquids 
include those of Nature such as elastically deform- 
able red blood cells suspended in plasma which may 
be reshaped with reduced diameters for easier flow 
through small blood vessels. 

In suspension applications, rheological fluids 
are attractive and their behaviour is represented in a 
simple way by Fig 5: constituent particles which are 
randomly distributed in a zero-energy field, become 
aligned with the field under excitation. Considerable 
changes in shear force between the plates occurs 
during excitation by the field, hence viscosity is 
controllable. Both electro- and magneto-field ex- 
cited types have been studied for many years. 

So-called electro-rheological magnetic (ERM) 
fluids can develop very high levels, above 50 psi, of 
fluid shear stress while operating on vehicle battery 

In controlling the fluid, the output signal from 
the ECU is transformed from an analogue voltage 
into a frequency, in the V/F converter, and used in a 
pulse-width-modulated power driver causing a cur- 
rent level to flow to a set of electro-magnets. Fig 8 
shows a schematic of the ECU, the left side depicting 


MonZero Field 

Porticlej Align In 
Column Format ion 

Fig 5: Particle orientation in rheological fluid 

voltage levels or moderate currents. This fluid also 
has the ability to set up shear forces independent of 
velocity, so no fluid flow is required. With such a 
fluid, a variable suspension damping characteristic is 
possible while reverting from current, space-wasting 
types to the more compact rotary configurations of 
earlier years. A typical parallel lower control arm 
suspension is shown in Fig 6. The electronic control 
allows an algorithm to process information for the 

Fig 8: Electronic control unit 



the inputs from the sensors which are converted for 
processing by the DSP and then reconverted to ana- 
logue for transformation to the appropriate pulse- 
width-modified signal. The ECU can tune the vehi- 
cle through software without any hardware change. 
This could involve mere days of work compared with 
the months currently involved in tuning the dampers 
for a new vehicle platform. 


The multi-degree of freedom matrix equation 
to control the dampers according to both vehicle and 
road dynamics is represented by the algorithm of Fig 
9. For the future, more complex algorithms and 
integrated electronics are forecast by the authors. 
Microprocessor instructions could be executed at the 
rate of millions per second, transforms between time 
and frequency domains becoming possible and hence 
more complex control strategies permitted. 

Fig 7: Cross section of RACD damper unit 

Electronic control 
of electric steering 

Following Honda’s successful pioneering of electric 
steering on the NSX car, company engineers 4 have 
reported a new-generation system for future applica- 
tion of intelligent steering. Fig 10. Its lighter and 
more compact construction allows its substitution for 
hydraulic PAS on a wide range of light vehicles. 

The torque sensor, Fig 11, on the steering 
column relates steering input torque from the driver 

Fig 9: ERM/RACD control algorithm 



with axial displacement and senses change of induct- 
ance in its field coil. The steering servo motor driving 
the rack has current feedback control and as well as 
the input from the column torque sensor also has a 
vehicle input related to instantaneous road speed; a 
PWM drive with power MOSFET is involved. 

The seamless design of the torque sensor, inte- 
grated into input shaft, allows driver input and road 
kick-back to be evaluated. On the input shaft, steer- 
ing wheel turning speed is derived from differential 
value of the torque as well as the direct value. On the 
output shaft, motor assist-torque data, and the reac- 
tion torque from the road, is obtained as a micro- 
vibration trace. Information on the inertia of the rotor 
is also obtained from the output shaft and the detri- 
mental increment of inertia is cancelled out electri- 
cally to prevent its transmission to the steering wheel. 

In the ECU for the new generation system, the 
microprocessor performance has been raised from 8 
bit 12 MHz to 16 bit 16 Mhz. ASICs are now 
employed to customize the input and output circuits. 

with a resultant 35% reduction in part-count. Be- 
cause the steering motor requires some 10 amperes 
for control, MOSFET switching and metal strip 
conductors are exploited with specially developed 
connectors. Motor and control-unit are now adjacent 
to one another so as to minimize cable losses. 

Input shaft Torsion bar 

Information from Input shaft side 

• Input steering torque 

• Rotating speed at start of steering handle 

Pinion gear 

Structure o< torque sensor 

Information from pinion shaft side 

* Assist torque by motor 

* Resistance torque from road surface | 

* Kick back torque 

* Vibration of tire umbalance 

* Self-alignment torque 

Fig 11: Torque sensor Torques sensed by torque sensor 

EPS Control Unit 
{Integrated with PDU) 

Steering Column 

Torque Sensor 

Indicator light 

j 1 ^ inaitaiur 1 

“L I Constant voltage circuit Torque check drcu n 

Fig 10: Honda intelligent steering system 

Current sensor 



Drivetrain control 

The farther future in mechatronics is seen by re- 
searchers at Vairex Corporation 5 as the electric 
drivetrain. Full drive-by-wire will eliminate me- 
chanical connections and replace them by electri- 
cally powered components and electronic controls, 
with the eventual reality of each wheel being capable 
of instantaneously, accurately and independently 

absorbing or developing torque between tyre and 
road with positive or negative value on any surface — 
with a dramatic improvement in traction, braking and 
handling performance. A particular advantage would 
be the introduction of electro-magnetic braking of a 
type that is more efficient than regenerative systems 
being pioneered on electric and hybrid propulsion 
vehicles. With regen braking, maximum braking 

Conventional Drivetrain 

Electric Drivetrain 

Fig 12: 


(left) and 






150 kW 1C Engine 


50 to 75 kW IC Engine 


Automatic Transmission 


75 kW Alternator 


Logic Module 


Power Module 


Logic Module 


Transfer Case 




35 to 50 kW Wheel Motors 




CV Joints 


Disc Brake Assemblies 


ABS/SC Hydraulic Module 


Brake Hydraulic Pump/Accumulator 


Power Steering Rack 


Steering Rack 


Power Steering Pump 


40 L Fuel Tank 


20 L Fuel Tank 


Batterv/Capacitor Module 



force is limited by the rate at which the energy storage 
system can absorb and dispense power; without the 
use of ultracapacitors, therefore, only supplementary 
braking is possible electro-magnetically. Also brak- 
ing effectiveness is reduced to the degree that to the 
degree that zero braking force is reached slightly 
before the vehicle reaches zero speed. If reverse 
power rather than a load is applied to the generator, 
braking force down to zero speed can be generated 
and held there. The braking requirements of an 
average saloon car, say the authors, would require 
wheel motors, on each of the four wheels, with 
individual capacity of 35-50 kW. Fig 12 shows the 
author’s electric drivetrain proposal alongside a con- 
ventional drivetrain. This would suit a 1750 kg 
vehicle, having all-wheel-drive, electronic torque 
split, traction control, automatic transmission, ABS 
braking and stability control. For safety, a conven- 
tional friction park-brake would probably be re- 

Navigation system advances 

The 1998 SAE Congress in Detroit was the occasion 
for the reporting of important telematics develop- 
ments bringing closer the prospects of intelligent 
transportation systems on public highways. 

Researchers from Hitachi 6 described an adap- 
tive cruise control (ACC) system using a wheel 
torque management technique and involving a longi- 


Speed v c 

" ltT,s : Power train 

G^s| - — ^ : Approximated running resistance 


m : Vehicle mass 

distance D„ 

distance 0„ 

(Relative speed v , -sDj Fig 14: Speed controller ( top) 

tudinal control method. An electronically controlled 
throttle valve, electrically controlled brake and auto- 
matic transmission work co-operatively to provide 
smoother acceleration, better shift performance and 
better fuel economy in ACC mode (at a desired 
headway distance detected by a micro-wave radar 
sensor), while, in manual driving, wheel torque is 
proportional to throttle pedal displacement. 

The management system seen in the lower part 
of Fig 13 selects signals from throttle-control, trans- 
mission and brake to send commands which depend 
upon the requested wheel torque. Conventional cruise 
control is obtained by adding a speed controller to the 
wheel torque manager, and on top of this, headway 
control computes a command speed and sends it to 
the speed controller which calculates the required 
wheel torque. Block diagrams for the speed and 
headway controllers are shown in Fig 14. 

Engine " " u “ n “ ,c 


Fig 13: Constituents of ACC 

0 t, io 20 t z 30 t 3 40 50 

Time [sec] 

and headway controller Fig 15: Experimental results of ACC 



Tests carried out with the system installed in a 
vehicle showed the ACC performance as in Fig 15. 
From ACC start-up at t = 0, the preceding vehicle 
increases its speed v at time f ; then decreases its 
speed at time t } by braking and at time t 4 its speed is 
kept constant. The wheel torque shows its peak at the 
beginning of acceleration and keeps it a constant as 
speed increases; in the deceleration phase, at about 
-2.5 m/s 2 , the torque becomes negative under electri- 
cal braking. 

Headway sensor design 

The development of a suitable radar sensor for ACC 
has been described by engineers at Bosch 7 which is 
due for European introduction in the current model 
year in its initial form as a high-cost option on top-of- 
the-range vehicles. For more general application a 
less expensive version is under development, the first 
stage of which is an integrated sensor and control unit 
described in the paper. This would be incorporated in 
a typical ACC system depicted in Fig 16. 

The main functional advantage of integration 
is improved communication between sensor and 
controller and that not only the prime target preced- 
ing vehicle is considered but up to 8 objects in the 
path of the vehicle (valuable when a vehicle ahead of 
the preceding vehicle brakes suddenly). Basic speci- 
fication of the sensor is as follows: 



Relative speed 

-60m/s ... +60m/s 



Resolution cell 

0.6m , 1.7m/s 

Measurement rate 

10 Hz 

Freeways and inter-state highways with a mini- 
mum curve radius of about 500 m require a lateral 
detection angle as shown. A monostatic three beam 
frequency modulation continuous wave (FMCW)- 
type radar is employed, working at 76-77 GFlz, all 
three beams sending and receiving simultaneously. 
Characteristic data for the radar are: 

Frequency allocation 

76 - 77.0 GHz 

Total average output power 

< ImW 


220 MHz 

Number of beams 


Beam width (measured at the 
-6dB point of the radar- 


Overlap of two beams 
(„Squint Angle") 


Total angular range 

+/- 5° 

A block diagram for the integrated system is 
seen in Fig 17. 

Road guidance for drowsy drivers 

To provide vehicles with a guidance means based on 
white road-line recognition, a system based on lane 
region extraction and line edge detection has been 
described by Matsushita engineers. The proposed 
method uses both low and high spatial frequency 
image data, respectively, to detect these two at- 
tributes of the white lines. The company has shown 
that detection from 5 to 40 metres in front of the 
vehicle is accurately performed at a detection rate of 
over 99.4%. As well as protecting drivers who be- 
come drowsy the system also is used to distinguish 

Fig 16: ACC system 



present lane from adjacent lanes in automatic cruise 
control. The two separate modes of image detection 
described above are intended to overcome the prob- 
lems of intermittent white line fading and the prob- 
lem of the detection camera vibrating in sympathy 
with the vehicle. Fig 18 distinguishes between the 
two modes. Lane region extraction uses a knowl- 
edge-based brightness rule to distinguish the inside 
lane area and/or a low spatial frequency image in 
order to smooth the brightness of the lane region. The 
system checks there is a smooth transition between 
the dark area in front of the vehicle and the bright area 
further away. Edge data are determined using each 
threshold in each scan direction between 50 lines and 
200 lines, the following contour points being re- 
moved: those which are more than 20 dots away from 
the end-of-lane region (extracted from the low-fre- 
quency image); all non-continuous points and those 
which do not fit within the road width. Resultant 

white lines are calculated using a spline of 3 degrees, 
the detection process illustrated being implemented 
on a network of transputers. Camera-angle error 
compensation is based on distance between the lines 
being constant, contour points of lines in the image 
being transferred to points on the world co-ordinate 
mapped image seen in Fig 19. 

Lateral distance 

Fig 19: Mapped image 

Fig 17: Block diagram of 

Fig 18: 

White lines detection process 



Digital circuits for computation 

Logic gates (switching circuits) are known as build- 
ing blocks for electronic circuits used in computer 
control. A recent work 8 on engineering mathematics 
provides an unusually accessible description of the 
building of such circuits, in the author’s chapters on 
Boolean Algebra and Digital Systems. 

The Boolean variable, he explains, involves 
the realization of the two logical states, 0 and 1, 
analogous to the open and closed positions of a 
mechanical switch but in digital electronics repre- 
sented by low and high voltages in ‘positive’ logic. 
The logic gates. Fig 20, provide Boolean functions 
analogous to multiplication, division, inversion and 
their different combinations. 

Further combination of the basic gates may be 
used to form functional logic circuits such as that in 
Fig 21a which might be used for a vehicle warning 
buzzer so that it sounds when the key is in the ignition 
(A) and a door open (B) or when the headlamps are 
on (C) and a door open (A). Here there are two AND 
gates and an OR gate. The first two give the output 
A.B + C. A which is the input to the OR gate which has 
output A.(B + C). 

Fig 21a: Warning buzzer circuit 

Fig 20: Logic circuits, a. AND; b. OR; c. NOT; d. XOR; e. NAND; f. NOR 



In multiple combinations they come as rela- 
tively inexpensive integrated circuits, Fig 21b illus- 
trating the 7408 and 7402 circuits. The first has four 
two-input AND gates supplied as a 14 pin package, 
with power supply connected to pins 7 and 14, 
supplying the operating voltage for all the four AND 
gates. A notch is cut between 1 and 14 to show at 
which end of the package pin 1 starts. In the 7402 
package there are four two-input NOR gates. The use 
of 7408 as a clock enable circuit is shown in the view 
in Fig 21c, where signals are passed on to a receiving 
device only when a switch is set to an enable position. 
The clock emits a sequence of 0 and 1 signals 
(pulses), Fig 22. The AND gate will give an output 
when both input signals are 1. 

Computer control usually involves implemen- 
tation of arithmetic operations for which the binary 
numbering system is used having binary digits (bits) 
0 and 1 . Fig 23a shows a half-adder circuit giving the 
sum equal to 1 when either first or second input 
numbers, but not both numbers, is 1, obtained with an 
XOR gate. The carry-out bit is to be 1 when first and 
second numbers are both 1, obtained with an AND 

gate. A full adder circuit is seen in Fig 23b; the sum 
output arriving from the three inputs, first, second 
and carry-in number. This provides a 1 output when 
there is an odd number of 1 inputs, obtained by two 
XOR gates. The carry-out output is 1 when any two 
of the inputs are 1 , obtained with three AND gates and 
an OR gate. 

The half and full adder circuits can be com- 
bined to realize multi-bit adder circuits. Thus, for a 4- 
bit adder. Fig 24 is an appropriate circuit, in which the 
full and half adders are represented as blocks for 
simplicity. By connecting the carry-out from one 4- 
bit adder to the carry-in for another one, an 8-bit 
adder circuit can be obtained (used for adding posi- 
tive numbers together), Bolton explains. 





Fig 21c: Clock enable circuit using 7406 1C Fig 23a: Half-adder 

Sum number S^SjSjS^ 

Fig 24: 4-bit adder 

Fig 23b: Full adder 



Fig 25: Personal Productivity info-technologies 

Fig 26: TEC mobility system 

Proprietary control 
system advances 

The Delphi Personal Productivity vehicle is a com- 
bination of information technologies, Fig 25, for per- 
mitting business drivers to remain in constant com- 
munication with their offices. First demonstrated in 
a Saab car it has involved co-operation between the 
suppliers and Microsoft/Mecel. By means of Win- 
dows CE, installed in an on-board computer, the sys- 
tem links with office-based Windows PCs. A speech 
interface called Auto PC allows hands-free commu- 
nication for the driver with voice commands being 
used to send E-mail messages, for example. The fa- 
cility also uses speech synthesis to convert incoming 
messages for the driver. There is, too, the option of 
using steering-wheel mounted controls instead of 
audio communication. The Delco Mobile Media Link 
employed is a high-speed fibre-optic serial data link 
capable of transmitting data at 98 MB/sec. It also 
has sufficient band width to allow reception of up to 
50 channels of stereo audio and 1 5 channels of TV- 
quality compressed video. A cellular modem connec- 
tion allows information like traffic updates to be in- 
put from an internet provider. 

Following the introduction of the Magneti- 
Marelli Route Planner in-car navigation system, the 
next generation TEC mobility system under develop- 
ment, Fig 26, will allow the driver to access Pentium 
computer power through ConnectCar-PC via a voice 


MED 7 



recognition interface. As well as helping the driver to 
navigate, he/she will be able to place and receive 
telephone calls and even browse the internet or send 
an E-mail. 

The Bosch Motronic MED 7 engine manage- 
ment system, Fig 27, is designed for direct-injection 
gasoline engines and calls on a wide variety of sensor 
inputs before computing engine torque response to 
throttle-pedal displacement. Some 15-20% fuel sav- 
ings are claimed for DI engines using the system 
compared with conventional state-of-the-art gaso- 
line engines. 

Delphi’s E-STEER electronic steering incor- 
porates a steering gear assist mechanism , Fig 28, and 
electronic controller to provide responsive steering 
assist, in several configurations, Fig 29. Sensors 
measure two primary inputs — vehicle speed and 
driver torque (or effort) on the steering shaft at the 
assist unit. These two primary inputs, along with 
other system variables and inputs, are continuously 
fed into an electronic control module, which analyses 
the data and determines the direction and the amount 
of steering assist required. The controller then gen- 
erates a command to the variable-speed, 12-volt 
electric motor. Using intricate control algorithms, 
the controller’s command varies the torque of the 
electric motor, which drives a gear mechanism to 
provide the required assist. A permanent magnet 
brushless motor is used, which allows the smallest 
package size, the smallest mass, and the lowest rotor 
inertia (for more responsiveness). The motor is also 
quieter, more reliable, and more powerful than many 
types of electric motors; it uses the vehicle’s battery 

as its power source. This can be especially important 
to drivers who experience an engine stall while 
turning a sharp comer or rounding a steep mountain 

Traditional hydraulic systems can be sluggish 
until fluids warm up, especially in very cold climates. 
The Delphi system provides cold-weather rapid start. 
With its engine independence and fluid-free design, 
the system is less sensitive to cold temperatures even 

Fig 29: Different configurations 



Fig 30a: Vehicle fuel consumption rating — comparison between manual steering and E-Steer; 




at -40C The system also eliminates the large energy 
loss typically involved in those first few minutes of 
warm-up. When a major European manufacturer 
conducted a fuel economy test, it found virtually no 
difference in fuel usage between this EPS and manual 
steering. Fig 30a. 

This EPS is an ‘on-demand’ system, running 
in a monitoring mode until assist is needed. Typically 
using less than 0.5 amp at idle and with the average 
usage between 1 to 2 amp, the system has extremely 
low parasitic energy losses for maximum efficiency. 
In addition to reducing engine drain, the system may 
also reduce overall mass. The system also provides 
improved stability for greater passenger comfort, 
control, and safety. There is smoother transition from 
one effort level to another — without compromising 
power steering assist between parking manoeuvres 
and highway driving . 

System design with more options Speed-sensi- 
tive variable effort steering is an increasingly popular 
option, but the need for a controller makes it a costly 
addition to conventional hydraulic systems. Delphi’s 
E-Steer can calculate the required assist level, based 
on vehicle speed and driver torque input, using its 
existing software and controller, Fig 30b. This makes 
speed-sensitive variable effort steering (and other 
options) available for very little added cost. 

A further benefit of electric power steering is 
that assist can be provided in both directions, so it is 
no longer necessary to compromise suspension de- 
sign in order to achieve the required retumability 
characteristics. The system’s optional Assisted Re- 
turn feature allows the retumability to be optimized 
independently, so that when the driver turns and then 
releases the wheel, E-Steer returns it safely to centre. 
The company has also developed a unique Steer 

Damping option that helps to provide a smoother feel 
and reduce the effects of issues such as free control. 
A conventional hydraulic steering system would 
require time-consuming and expensive mechanical 
modifications, but with E-Steer this can be achieved 
with software. 

Fast system tuning The operating characteris- 
tics of any steering assist system must be carefully 
tuned to match the vehicle design objectives. With 
conventional hydraulic systems, tuning requires the 
removal of the steering gear and the manufacture and 
fitment of modified valves and torsion bars. This 
process limits testing to two or three configurations 
per day, with the whole process typically taking from 
one to six months. E-Steer allows very fast tuning 
without mechanical intervention, allowing the de- 
sired feel and performance to be achieved in a small 
fraction of the time previously required. All that is 
required is a laptop computer with Delphi’s propri- 
etary tuning package. 

Reliability In addition to engine independ- 
ence, there are specific reliability features built into 
the system. During development, company engi- 
neers set up testing situations to simulate possible 
faults in the system. With this knowledge, they built 
in several reliability features. The system continu- 
ously runs detailed self-checks and diagnostics, en- 
suring that all areas of the system are functioning 
properly. Redundancy (back-up or ‘repeat’ system 
elements) has been engineered into the system to 
provide reliability. 

Packaging flexibility With no pumps, hoses, or 
belts on the engine, this simplified packaging pro- 
vides maximum design and engineering flexibility. 
The system controller can be incorporated with the 
column, the rack, the pinion, or in a stand-alone 

oNivm Tonoua 

Fig 30b: Performance parameters: left: steer damping; centre: tuning capability; right: variable effort 



location. The controller can be in the engine or the 
passenger compartment. The system frees up the part 
of the engine where a belt and pulley to drive the 
pump would otherwise be placed. This makes room 
for other features like pollution control or an air- 
conditioning compressor. And there are no more 
hoses snaking through the engine compartment. The 
system can help reduce the number of parts the car 
manufacturer has to track, order, and store. For 
example, two models in the same platform, a 4-door 
sedan and a 2-door coupe, can use the same physical 
steering system. With the advanced tuning technol- 
ogy of the system, one could custom tune a luxury 
feel for the sedan and a sporty feel for the coupe 

For the future Delphi Steering anticipates the 
development of several options for future electric 
power steering systems. Potential combinations in- 
clude selectable steering to go with a selectable 
chassis where the driver could push a button for a 
sportier feel, for example. Hand-held diagnostic tools 
could be used at the dealership, allowing consumers 
to request changes to the feel of their steering. As a 
cost-reduction strategy, the system could be very 
useful in advanced systems integration. With up- 
level integration, algorithms/software for other sys- 
tems could be combined — like security, collision 
avoidance, steer by wire, or automated highway 

Fig 32: Use of vehicle per day 

Hybrid drive prospects 

A neat description of the problems of hybrid drive 
vehicles has come out of the results of the three-year 
HYZEM research programme undertaken by Euro- 
pean manufacturers. According to Rover partici- 
pants 9 , controlled comparisons of different hybrid 
drive configurations, using verified simulation tools, 
are able to highlight the profitable fields of develop- 
ment needed to arrive at a fully competitive hybrid 
drive vehicle and demonstrate, in quantitative terms, 

Electric Energy Consumption (kWh/1 00km) 
Independent hybrid' Substitution hybrid^ 

Fig 31: Characterizing a hybrid powertrain 

Fig 34: Synthetic urban drive cycle 

Fig 33: Daily distances 
and trip lengths 

Days Number in % 

o Athens 
o Germany 
A France 


0 20 40 60 80 100 150 200 Last 

Daily Distance travelled (Vm) 



the trade-off between emissions, electrical energy 
and fuel consumption. Only two standard test points 
are required to describe the almost linear relation- 
ship: fuel consumption at point of no overall change 
in battery state-of-charge (SOC) and point of electri- 
cal consumption over the same cycle in pure electric 
mode. A linear characteristic representing an ideal 

Central Battery Electric 

Fig 35: BMW parallel hybrid drive 

Fig 36: Parallel hybrid drive mechanism 

Fig 38: Ragone diagram for the two battery systems 

lossless battery can also be added to the graph, to 
show the potential for battery development, Fig 31. 

Confirmation was also given to such empirical 
assessments that parallel hybrids give particularly 
good fuel economy because of the inherent effi- 
ciency of transferring energy direct to the wheels as 
against the series hybrids’ relatively inefficient en- 
ergy conversion from mechanical to electrical drive. 
The need for a battery which can cope with much 
more frequent charge-discharge cycles than one for a 
pure electric-drive vehicle was also confirmed. Al- 
though electric energy capability requirement is less 
stringent, a need to reduce weight is paramount in 
overcoming the problem of the redundant drive in 
hybrid designs. 

A useful analyis of over 10,000 car journeys 
throughout Europe was undertaken for a better un- 
derstanding of ‘mission profile’ for the driving cycles 
involved. Cars were found to be used typically be- 
tween 1 and 8 times per day, Fig 32, and total daily 
distances travelled were mostly less than 55 km. 
Some 13% of trips. Fig 3, were less than 500 metres, 
showing that we are in danger of becoming like 
Americans who drive even to visit their next door 
neighbours! Even more useful velocity and accelera- 
tion profiles were obtained, by data recoding at 1 Hz 
frequency, so that valuable synthetic drive cycles 
were obtained such as the urban driving one shown 
in Fig 34. 

51 8i Hybrid 

51 8i standard 

Combustion engine 

1 ,8-ltr. 4-cyl 

162 Nm (119 Ib-tl). 
83 kW (113 Dhp) 

162 Nm (119 Ib-ft) 
83 kW (113 bhp) 

Electric motor/ 

Siemens asynchronous 
200V. max 280A, 95 Nm 
(70 Ib-ft) 

(165 Nm/122 Ib-ft max). 
18 kW (26 kW max) 

Transmission ratios 

5 09/2 8/1 76/1 24/1.0 

5.1/2 77/1.72/ 

Final drive 


3 46 

Ni/MH battery 

Cell type: X40. 180 1.2V 
cells/40Ah/1.06 kg. total 
energy content 9 kWh. 
Dimensions: 803/465/ 
418 mm 3 

Ni/Cd battery 

DAUG-Hoppecke. cell 
type: X35, 168 1.2V 
cells/35Ah/1 .15 kg 
Energy content 7 kWh, 
Dimensions: 859/400/410 mm 3 

Weight, unladen kg (lb) 

1940 (4278) 


Performance weight 
kg (lb) 


1618 (3568) 

Fig 37: Vehicle specification 



BMW researchers 10 have shown the possibility 
of challenging the fuel consumption levels of con- 
ventional cars with parallel hybrid levels, by using 
map-controlled drive management. The 2-shaft sys- 
tem used by the company, Fig 35, uses a rod-shaped 
asynchronous motor, by Siemens, fitted parallel to 
the crankshaft beneath the intake manifold of the 4- 
cylinder engine, driving the tooth-belt drive system 
as seen in Fig 36; overall specification compared 
with the 5 1 8i production car from which it is derived 
is shown in Fig 37. The vehicle still has top speed of 
1 80 kph ( 1 00 kph in electric mode) and a range of 500 
km; relative performance of the battery options is 
shown by Fig 38. Electric servo pumps for steering 
and braking systems are specified for the hybrid 
vehicle and a cooling system for the electric motor is 
incorporated. The motor is energized by the battery 
via a 13.8 V/ 50 A DC/DC converter. The key 
electronic control unit links with the main systems of 
the vehicle as seen in Fig 39. 

To implement the driving modes of either 
hybrid, electric or IC engine the operating stategy is 
broken down into tasks processed parallel to one 
another by the CPU, to control and monitor engine, 
motor, battery and electric clutch. The mode task 
determines which traction condition is appropriate, 
balancing the inputs from the power sources; the 
performance/output task controls power flow within 
the total system; the battery task controls battery 
charging. According to accelerator/braking pedal 

Clutch System 


[ Digital Motor 
! Electronics 


> t s 








..^Control Unit^ 







Brake Pedal 

Pedal Sensor 



Fig 39: Vehicle management 

Generator rpm Engine rpm Motor rpm 

The three vertical lines in the diagram shov the 
shahs in the planetar)- 

Fig 41: Power interaction diagram 

Fig 40: 

Efficiency h 



inputs, the monitoring unit transfers the power target 
required by the driver to the CPU where optimal 
operating point for both drive units is calculated in a 
continuous, iterative process. Fig 40 gives an exam- 
ple of three iterations for charge efficiency, also 
determined by the CPU, based on current charge 
level of the battery. 

Production control-system 

In the Toyota Prius production hybrid-drive car, 
described in a separate article in this issue, for the 
power-split device which is a key part of the system, 
company engineers" have provided the diagram of 
Fig 41 to show how the engine, generator and motor 
operate under different conditions. At A level with 
vehicle at rest, engine, generator and motor are also 
at rest; on engine start-up the generator produces 

electricity acting as a starter to start the engine as well 
as operating the motor causing the vehicle to move 
off as at B. For normal driving the engine supplies 
enough power and there is no need for generation of 
electricity, C. As the vehicle accelerates from the 
cruise condition, generator output increases and the 
motor sends extra power to the drive shaft for assist- 
ing acceleration, D. The system can change engine 
speed by controlling generator speed; some of the 
engine output goes to the motor via the generator as 
extra acceleration power and there is no need for a 
conventional transmission. 

The control system schematic for the vehicle is 
in Fig 42, the THS (Toyota Hybrid System) calcu- 
lates desired and existing operating conditions and 
controls the vehicle systems accordingly, in real 
time. The ECU keeps the engine operating in a 
predetermined high torque to maximize fuel economy. 
The corresponding schematic for the ECU is in Fig 
43. It is made up of 5 separate ECUs for the major 
vehicle systems. The hybrid ECU controls overall 
drive force by calculating engine output, motor torque 
and generator drive torque, based on accelerator and 
shift position. Request values sent out are received by 
other ECUs; the motor one controls the generator 
inverters to output a 3-phase DC current for desired 
torque; the engine ECU controls the electronic 
throttle in accordance with requested output; the 
braking ECU co-ordinates braking effort of motor- 
regeneration and mechanical brakes; the battery ECU 
controls charge rate. 

Fig 42: THS control system 

- Anion Cation 

(Glass fiber) 

Fig 44: EDLC model (a), above, and cell (b), below Reduction gear* 

Electricity path 
Power path 

Fig 43: ECU schematic 



Energy storage initiatives 

According to researchers at NEC Corp 12 , the super- 
capacitor will be an important contributor to the 
energy efficient hybrid vehicle, the absence of chemi- 
cal reaction allowing a durable means of obtaining 
high energy charge/discharge cycles. Tests have 
shown for multi-stop vehicle operations a 25-30% 
fuel saving was obtained in a compact hybrid vehicle 
fitted with regenerative braking. 

While energy density of existing, non-automo- 
tive, supercapacitors is only aboutl0% of that of 
lead-acid batteries, the authors explain, it is still 
possible to compensate for some of the weak points 
of conventional batteries. For effective power assist 
in hybrids, supercapacitors need working voltage of 
over 100 V, alongside low equivalent series resist- 
ance and high energy density. The authors have 
produced 120V units operating at 24 kW fabricated 
from newly developed activated carbon/carbon com- 
posites. Electric double layer capacitors (EDLCs) 
depend on the layering between electrode surface 
and electrolyte, Fig 44(a) shows an EDLC model. 
Because energy is stored in physical adsorption/ 
desorption of ions, without chemical reaction, good 
life is obtained. The active carbon electrodes usually 
have specific surface area over 1000 m 2 /g and dou- 
ble-layer capacitance is some 20-30 (iF/cm 2 (acti- 
vated carbon has capacitance over 200-300 F/g). The 
EDLC has two double-layers in series, so it is possi- 
ble to obtain 50-70F using a gram of activated 
carbon. Working voltage is about 1.2V and storable 
energy is thus 50 J/g or 14 Wh/kg. 

Fig 44(b) shows a cell cross-section, the con- 
ductive rubber having 0.2 S/cm conductivity and 
thickness of 20 microns. The sulphuric acid electro- 
lyte has conductivity of 0.7 S/cm. Fig 45(a) shows the 

Fig 46a: Constant power discharge characteristics 



Working voltage (V) 


Capacitance (F) 


ESR (mil) 


Maximum current (A) 


Weighi (kg) 


Volume (1.) 


Size ( W x D x II mm) 

390 x 270 x 160 

Power density (kW/kg) 


Power density (kW/L) 


Energy density < Wh/kg) 


Energy density (Wh/L) 


Fig 45: High-power EDLC schematic ( above )and 
specification (below) 


0.1 1 0 10 100 1000 
Energy density / Wh kg - 1 

Fig 46b: Power density, y-axis in W/kg, vs energy 
density for high-power EDLC 

Accelerator sensor (High-po wer E DLCs) 
Power module 




ECU \ AC generator 

Lead-acid battery 

Fig 47: ELCAPA configuration 



high-power EDLC suitable for a hybrid vehicle and 
the table in Fig 45(b), its specification. Plate size is 
68 X 48X1 mm 3 and the weight 2.5 g, a pair having 
300 F capacity. Fig 46(a) shows constant power 
discharge characteristics and Fig 46(b) compares the 
EDLC’s energy density with that of other batteries. 

Fuji Industries’s ELCAPA hybrid vehicle, Fig 
47, uses two EDLCs (of 40 F total capacity) in 
parallel with lead-acid batteries. The stored energy 
can accelerate the vehicle to 50 kph in a few seconds 
and energy is recharged during regenerative braking. 
When high energy batteries are used alongside the 
supercapacitors, the authors predict that full com- 
petitive road performance will be obtainable. 

A high energy battery receiving considerable 
attention is the lithium ion cell unit, the development 
of which has been described by Nissan and Sony 
engineers 13 who point out that because of the high cell 
voltage, relatively few cells are required and better 
battery management is thus obtained. Accurate de- 
tection of battery state-of-charge is possible based on 
voltage measurement. In the battery system devel- 
oped, Fig 48, cell controllers and a battery controller 
work together to calculate battery power, and re- 
maining capacity, and convey the results to the 
vehicle control unit. Charging current bypass circuits 
are also controlled on a cell to cell basis. Maximizing 
lifetime performance of an EV battery is seen by the 
authors to be as important as energy density level. 
Each module of the battery system has a thermistor to 
detect temperature and signal the controllers to acti- 
vate cooling fans as necessary. 

Nissan are reported to be launching the Ultra 
EV in 1999 with lithium-ion batteries; the car is said 

to return a 120 mile range per charge. For the farther 
future lithium-polymer batteries are reported to be 
capable of giving 300 mile ranges. 

There is also less strength in the arguments 
against EVs with news of cheaper solar cells. Rapid 
thermal processing (RTP) techniques are said to be 
halving the time normally taken to produce silicon 
solar cells, while retaining an 1 8% energy conversion 
efficiency from sunlight. Researchers at Georgia 
Institute of Technology have demonstrated RTP 
processing involving a 3 minute thermal diffusion, as 
against the current commercial process taking three 
hours. An EC study has also shown that mass-produc- 
tion of solar cells could bring substantial benefits and 
that a £350 million plant investment could produce 
enough panels to produce 500 MW annually and cut 
the generating cost from 64p/kWh to 13p. 

A move towards fuel cells for electrical power 
generation is also gaining ground. Ballard Genera- 
tion Systems has announced the start-up of a PEM 
fuel cell, natural gas fuelled, with 250 kW capacity 
and using Johnson Matthey electrodes and platinum 
catalysts. Compared with on-board fuel cells, the 
larger stationary generating ones can better exploit 
the high operating efficiency in relation to the rela- 
tively high capital cost. The US Dept of Energy 
argues that over one year 1 MW of electricity gener- 
ated by fuel cells will avoid 45 tonnes of sulphur 
dioxide and 1 9 tonnes of nitric oxides being emitted. 

Ballard are also in joint agreement with Daim- 
ler-Benz on in-vehicle fuel cells and an A-class 
Mercedes-Benz, thus powered, is scheduled for around 
2004. Mass production to bring down the cost of fuel 
cells is the objective. 



Automation of handling tests 

Recent publicity surrounding deaths of road users 
due to accidents caused by OEM’s road-test drivers, 
and the need to reproduce laboratory conditions of 
repeatability into track testing, will hasten the ac- 
ceptance of the ABD steering robot, announced by 
Anthony Best Dynamics. Essentially the device com- 
prises a motorized assembly (which replaces the 
steering wheel) whose stator is linked to a pair of load 
cells measuring torque reaction; these are, in turn, 
coupled to an adjustable arm clamped on to a pillar 
member between the vehicle floor and suction caps 
on the windscreen. 

The computer-controlled steering robot pro- 
vides steering inputs to vehicles during track testing. 
Historically, this type of testing has been carried out 
manually, but it has been difficult to ensure that 
accurate and repeatable inputs are generated. The 
system developed by ABD overcomes these prob- 
lems, making it easy to obtain meaningful compari- 
sons between different vehicles and vehicle configu- 
rations. The robot, whichcan be used in cars, vans and 
HGVs, is capable of performing the range of tests 
described in the international standard ISO7401 as 
well as many special tests, and has a specification 
bom out of detailed discussions with a number of 
potential customers who all had considerable experi- 
ence in this type of testing. The computer and control 
technology employed is a spin off from ABD’s 
successful Suspension Parameter Measurement Ma- 
chine (SPMM) installed in both the UK and the US. 
The SPMM is a cost-effective four wheel-station 
machine which can measure the suspension, kin- 
ematic and compliance characteristics of a wide 
range of vehicles. The new steering robot uses tech- 
nology transferred from the SPMM to apply steering 
inputs to vehicles so that their transient response 
characteristics can be studied, providing accurate 
and repeatable inputs even in low frequency sinusoi- 
dal tests. 

General objectives 

At the heart of the robot is a direct drive DC torque 
motor with incremental encoder feedback. This pro- 
vides the motive force in a small package which fits 
within the diameter of a standard car steering wheel. 
The steering robot therefore only requires a DC 
electrical supply, which is normally taken from the 

vehicle’s own electrical system. 

The steering robot gives very precise, smooth 
control of the steering without the need for gears or 
clutch, in a high performance package with mini- 
mum weight and inertia. The direct drive design also 
enables the driver to steer the vehicle easily when the 
unit is de-energized, as there is very little torque 
required to back-drive the motor. This enables the 
unit to be used to leam driver inputs, which can then 
be replayed by the motor, as well as providing a very 
simple and inherently safe construction. 

The stationary side of the motor is connected to 
the vehicle body via a special universal torque reac- 
tion mechanism utilizing a built-in torque trans- 
ducer, which is installed in the passenger side of the 
vehicle between the windscreen and the floor. Fig 49. 
The system is designed to be installed easily, to 
provide no access restrictions for the driver, to cause 
minimal loss of field of view and to be very rigid. The 
torque reaction system accommodates steering col- 
umn run out without generating side forces or degrad- 
ing the torque measurement. The motor assembly is 
normally fitted to the steering column in place of the 
vehicle’s standard steering wheel, as this provides a 
very compact installation with minimum inertia. 
However, an adaptor plate can be supplied to enable 
attachment to the rim of the standard steering wheel 
for situations where it is difficult to reproduce the 
splines on the steering column or where only a small 
amount of installation time is available. 

The steering robot is controlled by a portable 
industrial PC running a Windows NT based software 
package. The software enables a database of vehicles 
and tests to be built up with great flexibility in the test 
specifications. A range of sinusoidal tests are pro- 
vided as standard; single period, continuous and 
swept sine, as well as step and pulse inputs. Alterna- 
tively, test profiles can be played out from points 
stored on a disk file which have been generated 
elsewhere (special tests or random inputs for exam- 
ple) or learnt from driver inputs. During operation, 
tests are selected and initiated using a touch screen, 
which also provides status information as the test 
progresses. Whilst the robot is operating the driver 
holds a joystick which gives him the ability to adjust 
the steering drift of the vehicle and to attenuate the 
playout amplitude. The robot is automatically de- 
energized if the joystick is released to enable the 
driver to regain manual control of the steering. 



Fig 49 : Steering robot installed 

Fig 51: Computer test screen 

The computer system also has the ability to 
capture and store run-time data from the robot’s own 
transducers as well as up to twelve user-defined 
analogue inputs and three encoder inputs. The system 
also provides digital outputs to trigger external data 
capture, and analogue outputs of motor torque, angle 
and velocity. 

The new test systems ABD has developed are 
a response to an increased thrust in the industry to 
reduce time from concept to production of a new 
vehicle, with increased use of computer analysis and 
prediction techniques. Initial design parameters may 
be based on experience but, increasingly, use is being 
made of measurements on previous models and on 
competitor ‘target’ vehicles. Steering and suspen- 
sions are refined with the aid of models for which 
measured data is essential; also managers need to 
ensure that development work is carried out only on 
vehicles that are known to be built to specification. 
All this can be achieved only by making the relevant 

A particular specification 

The following specification describes a robotic ma- 
chine for applying low frequency inputs to a vehi- 
cle’s steering system to enable repeatable measure- 
ments to be made whilst testing on a track. Fig 50 
shows a schematic diagram of the robot. On the right 
of the figure is the servo-motor and angular position 
encoder which mounts directly to the vehicle’s steer- 
ing column. The torque reaction frame also functions 
as a stand for a flat, touch-screen, display (not 
shown). This is the user interface for the computer 
system, which is housed in a separate case located 
elsewhere in the vehicle (and not shown in the 
figure). A second, stacking case houses the motor 
amplifier and signal conditioning for the torque 
measuring system. Manual steering of the vehicle is 
still possible because the servo-motor is a direct- 
drive type and will back-drive easily when the power 
is removed. The motor has a steering wheel mounted 
around it to enable the vehicle to be driven normally. 
The motor is servo-controlled using position feed- 
back from an encoder. This task is undertaken by a 
dedicated motion controller. The user interface, com- 
munication with the motion controller, and other non 
time-critical functions are handled by a Pentium- 
based industrial computer. 

A telescopic pole between the vehicle floor and 
windscreen functions as the torque reaction frame for 
the motor. The two sections of the pole form a 
pneumatic actuator which is extended during fitting 
by using a low pressure air supply. Once firmly in 
position, a cross bar may be adjusted to allow the 
force links to be fitted to the drive unit at the correct 
angle. The two force links use charge devices to 
measure their axial loads. The steer torque is calcu- 
lated from these loads by the computer. The torque 
reaction frame also acts as a mount for a removable 
control box which houses the colour, TFT, display 
with a touch screen. This provides the principal 
control interface during testing on the track. Another 
control box, mounted to the driver’s window with 
suction pads, holds a two-axis joystick incorporating 
a button for activating closed-loop control of the 
motor. The spring centred left-right plane of the 
joystick adjusts the DC steering position, enabling 
the driver to cancel any drift of the vehicle during a 
test, while the forwards-backwards plane adjusts the 
attenuation of the test amplitude. The forwards- 
backwards plane has a built in friction brake to 



prevent accidental movement when altering the steer- 
ing position. A third control box on a flying lead 
contains an emergency stop button, to cut all power 
to the motor, and a button which can be used to 
initiate tests (as an alternative to using the touch 
screen). This box can be placed in any convenient 
position within the driver’s reach. The computer 
system, motor amplifier and signal conditioning hard- 
ware are housed in two separate rugged cases. These 
are designed to stack one on top of the other and have 
strong handles to enable them to be strapped securely 
within the vehicle. Interconnection cables between 
the boxes are arranged on one side of the boxes and 
are protected by hinged covers. The connections to 
the control devices, motor, transducers and power 
supply are all on the opposite side. The connectors 
are protected when not in use by removable lids. The 
control cases can be mounted at any convenient 
position, typically a front or rear seat. The main 
power supply is provided via a large DC to DC 
converter. This is typically located in the foot-well of 
the vehicle. 

The steer angle, and torque, are measured and 
recorded by the steering robot and output on +10V 
analogue (DAC) channels for sampling by other 
equipment. In addition, velocity or acceleration can 
be output on a third DAC channel. A further 12 
analogue input (ADC) channels and 3 encoder chan- 
nels are available for monitoring any other signals 
during a test. The scale factors (EU/V) and offsets for 
all these channels are adjustable to maximize the 
dynamic range. The software features a database 
which records all the vehicles that have been tested, 
and catalogues the tests performed and the results. 
Fig 51 shows a typical screen for selecting the test 
vehicle. The results files contain the steer angle, 
velocity and acceleration, the torque values, and the 
spare analogue and encoder channels all sampled at 
100 Hz. The files are available in ASCII format with 
a header describing the test and each channel of data. 
A facility to copy files to a network, or other drive, is 
included to facilitate transferring the results to other 
post-processing software. A number of standard tests, 



or ‘templates’, are available to make setting up tests 
for new vehicles easier. It is possible to add extra 
templates if required. 

During a test, the steer angle and test time 
remaining are displayed digitally on the touch screen 
in large text. The steer torque is also displayed but in 
the form of an analogue bar graph. In addition up to 
four of the spare channels can also be displayed, one 
digitally (primarily intended for a speed display if 
available), and three as bar graphs (for lateral accel- 
eration). A target limit can also be set for each spare 
channel. Reaching the target level during the test 
causes the colour of the display to change to give a 
clear visual indication to the driver. The bar graphs 
also have a peak hold facility to show the maximum 
levels achieved in each direction during the test. 

Performance modes 

The software performs the following types of steer- 
ing test: in the continuous sine test, for a fixed 
frequency sine output, the user can select the fre- 

quency, amplitude and test time. The playout is 
automatically ramped at start and finish to help 
maintain a straight vehicle heading. In the sine sweep 
test, the user can select the test time, start and finish 
frequencies and a start and finish amplitude. A facil- 
ity to generate a sweep with constant peak velocity is 
also available. In this mode the software calculates 
the instantaneous amplitude necessary to maintain 
the peak velocity constant throughout the test. The 
sweep is logarithmic in frequency. There is also a 
single period sine test for lane change testing. The 
amplitude, frequency and direction can be specified. 
Pulse tests, in addition, allow step or ramped inputs 
to be applied as a half cycle (move to an angle, dwell, 
move back to zero). The overall time, step angle and 
ramp rate can be specified. In User data test mode the 
software outputs data contained in a file. The soft- 
ware will check the file to make sure that the robot is 
capable of generating the speed required. The data 
will have previously been saved to the file by using 
the ‘leam mode’, or may be generated externally and 
imported from an ASCII file. The software also has 
a leam mode to record steering movements input by 
a driver and save the results in a file. The file can then 
be played out as a test 





Installation and operation 

A hole through the motor assembly makes attach- 
ment to the steering column easy, Fig 52. Installation 
of the torque reaction system is very simple and 
requires no tools other than a low pressure air supply. 
The computer system can be installed at any suitable 
place in the vehicle within 2 metres of the driver, and 
the display and joystick are easily attached. 

The database would normally be set-up for a 
new vehicle and any tests defined whilst the vehicle 
is in the running-shop. A normal PC keyboard (and 
mouse) would be used at this stage. Simple modifica- 
tions can be made to the test setups later while on the 
road, by using the touch screen. Once at the test site, 
the driver would set the correct direction and veloc- 
ity, and would press and hold the joystick button to 
engage the motor. The robot will then be in control of 
the steering. Any drift of the vehicle may be can- 
celled by the driver by moving the joystick in the left- 
right plane. Pressing the touch screen button (or 
remote button) will start the first test in the sequence, 
and cause a digital output to change state to signal to 
external data capture systems. The playout will begin 

after a preset delay (defined in the test setup). A 
second digital output will signal the start of the 
playout. During the test, the joystick up-down plane 
controls attenuation of the test amplitude. Releasing 
the joystick button at any time will disengage the 
motor and allow manual steering. Pressing the touch 
screen during a test will stop the test and reset the 
steering to straight ahead. Tests may be modified and 
repeated via the touch screen or by using the key- 

During the test the display will show: 

1) The steering wheel angle (large text) 

2) Direction of first turn in the test 

3) Warning that the robot is not following the test 
profile to a user-set tolerance 

4) Percentage test amplitude (100% clearly indi- 

5) Test time remaining 6) Test and robot status 
7) Bar graph displays of ADC and encoder channels 

The data from a series of tests would be saved 
to disk for post processing later, but results can be 
viewed graphically on the touch screen immediately 
a test is complete. 

Chapter 5: 

Vehicle development 

Among the recent vehicle introductions, four stand out as 
marking a major innovative input to vehicle development. 
The Mercedes A-class is a significant advance in small car 
packaging as well as breaking new ground in secondary safety. 
Ford’s Focus brings the company’s 21st century hard-edge 
styling to its most mainstream model and also marks a 
completed market repositioning of a class of vehicle starting 
as the Mark 1 Escort to a vehicle refined enough to compete 
in the neo-luxury sector. The Land Rover Freelander brings 
the price of comfortable 4-wheel drive vehicles down to 
realistic levels and is a breakthrough in body option versatility. 
Project Thrust SSC presents the ultimate in competition 
vehicles and gives a taste of the technologies required at the 
limits of vehicle-to-ground adhesion. 



Mercedes A-class 

The world’s pioneering motor manufacturer, and 
possibly still the most prestigious, had a momentous 
year in 1998, and seemed to emerge as the senior 
partner in Daimler-Chrysler. Its earlier attempts to 
break into compact car making have also now borne 
fruit in the re-launched moose-proof A-class, Fig 1, 
which has become possibly the class-leader in city 
cars. The revised suspension system is said to make 
the A-class the safest car in its class. The modified 
rear suspension, Fig 2, now has the special tie rod, 
shown, incorporated and all models benefit from 
ESP, the company’s Electronic StabilityProgramme. 
Chassis was also slightly lowered, anti-roll torque 
increased, rear track extended slightly, dampers/ 
tyres replaced, suspension stiffened and greater 
understeer built into the handling dynamics. 

Secondary safety systems 

The sandwich floor concept allows the engine and 
drive assembly of the A-class to be shunted back 
underneath the body in the event of a frontal impact 
rather than penetrating the passenger compartment. 
This ingenious layout, Fig 3, means that the new 
Mercedes model virtually achieves the high safety 
standard of the E-class. Even in the event of a side 
impact the ‘sandwich’ concept of the double layer 
floor, Fig 4, provides clear advantages, because the 
occupants sit about 20 centimetres higher than in 
other cars, with the result that the impact occurs 
below the occupant cell. The A-class complies not 
only with the future EU frontal impact crash test 
standard but also with strict safety regulations on side 
impacts in the USA and the European Union. In terms 
of safety a car measuring only 3.57 metres long with 
a kerb weight of only about 1000 kg has a number of 
basic disadvantages to overcome — especially if it is 
built on conventional lines. The low vehicle mass, 
short front-end structure and limited space available 
for crumple zones are fundamental handicaps that 
until now have left little room in this vehicle class for 
further progress in vehicle safety. This situation has 
been revolutionized with the A-class. 

The main aim in the safety development of the 
car was to make maximum use of the short front 
crumple zone in a head-on collision and to clear all 
components that obstructed the crumpling process 
out of the way. Because of the rigid structure of 
engine and gearbox, they play practically no part in 

T a 

Fig 1: Mercedes A-class 

Fig 2: Revised rear suspension 

Fig 3: Slant-mounted power unit 



the deformation, and in conventional compact cars 
they can be pushed back in a single block that can 
intrude into the interior and injure the occupants. 

Here the innovative arrangement of engine and 
gearbox reduces the risk of their forming a single 
block and therefore makes an essential contribution 
to occupant safety. In the A-class, the engine and 
gearbox are positioned at an angle partly in front of 
the occupant cell and partly underneath it, so that in 
the event of frontal impact they can slide downwards 
rather than straight back. This is possible due largely 

Fig 4: Double-level floor 

Fig 5: Vehicle chassis-frame construction 

Fig 6: Deformation zones 

to three unusual features: the unique inclined con- 
struction of the engine and gearbox, the shape and 
installation of which are specifically designed to 
accord with the crash principle of the A-class; the 
development of a frame-type integral support to 
accommodate the engine and gearbox (Fig 5), which 
in a crash can disengage at two of its eight fixing 
points thereby allowing the engine and gearbox to 
slide downwards; thirdly the design of the stable front 
floor panel (pedal floor) as a surface along which the 
engine and gearbox can slide in the event of a crash. 

This maximum possible deformation length 
available at the front end of the Mercedes A-class is 
due solely to the fact that the rigid engine and gearbox 
are displaced under the passenger cell in a serious 
head-on collision. Without this capacity for evasion 
and the ability of the engine and gearbox to ‘duck’ out 
of the deformation area, the front end of the A-class 
would have to be about 25 cm longer in order to give 
the occupants the same degree of protection. The fact 
that the short crumple zone also does not subject the 
occupants to severe stresses as a result of the higher 
deceleration values is due to the retaining systems 
specially matched to the front end structure. 

Other advantages of the new body concept are 
the straight side members. They absorb high defor- 
mation forces right from the start of body deforma- 
tion and thereby permit much more favourable decel- 
eration characteristics than in cars of conventional 
design. As a result, the occupants share in the decel- 
eration of the vehicle earlier and over a longer period, 
which reduces the stresses to which passengers are 
subjected. The triggering threshold of seat belt 
tensioners and airbags on the driver and front passen- 
ger side can also be more precisely defined thanks to 
the longitudinal rigidity of the body structure. The 
front end structure has deformation zones, Fig 6, on 
three levels. The impact energy in the front area of the 
A-class is widely distributed over three planes: the 
frame-type integral support, which serves to accom- 
modate the front axle, engine and gearbox; the straight 
side members leading to the aluminium and steel 
front module; the second upper side member plane on 
a level with the headlamps. 

Since most frontal impacts are offset crashes 
with unilateral stressing of the front end structure, 
there are strong transverse connections between the 
impact zones designed to ensure that the deformation 
areas on the opposite side, remote from the impact. 



are also involved. This produces an homogeneous 
front end structure which will prove effective in all 
types of frontal collision. The front wheels also have 
a role to play in energy absorption. They form an 
additional load path and in an impact are braced at 
the stable nodal points of A-pillar and sill. This effect 
is particularly important in the frequent front end 
crashes with overlap on one side — the so-called 
offset collisions. 

M-B engineers have matched the strength of 
the occupant cell to the high front end deformation 
force. The raised flat floor, which is welded from 
below to a grid-like support structure, gives the 
occupant cell a high degree of stability so that it is 
essentially not deformed either in a frontal impact or 
in an offset crash with 40 or 50 percent overlap. 
Airbags and belt tensioners and a polystyrene insert 
in the footwell, which form part of the standard 
equipment on all Mercedes cars, complete the list of 
measures to protect the occupants. 

From the outset the ‘Big versus Small’ aspect 
was a prime consideration in the safety development 
of the A-class. In a head-on collision involving two 
cars of differing mass a suitable distribution of the 
deformation load between the two vehicles must be 
achieved so that there is no danger particular to the 
occupants of the smaller car. For some time now 
Mercedes-Benz has been developing its cars accord- 
ing to this principle and claims to be the first to test 
cars through crash tests against a deformable barrier. 
This test method, which will not be specified for all 
new cars in the European Union until October 1998, 

permits an especially realistic simulation of typical 
accidents involving oncoming traffic, and specifies a 
head-on collision at 56 km/h and 40 percent overlap 
against a deformable barrier. The A-class passes this 
test even at an impact speed of 65 km/h, correspond- 
ing to a significantly high stress loading. 

Compatibility means that in a collision involv- 
ing two vehicles they should activate one another’s 
crumple zones and be capable of transmitting the 
deformation load uniformly to both bodies. For this 
to happen it is necessary for the front end deformation 
resistance of large and small cars to approximate to 
one another. As an example, in the E-class the front 
end structure of this Mercedes model is constructed 
so that in a crash the impact energy for the other 
vehicle is reduced - — intelligent safety engineering 
means mutual protection. In a smaller car, however, 
the design principle is just the opposite. Because of 
the lower mass and the shorter front end the Mercedes 
engineers designed the A-class with an especially 
rigid front end so that in an accident with another, 
larger oncoming vehicle it can activate the other 
vehicle’s deformation zones. In addition to matching 
the levels of force, the uniform distribution of impact 
forces over the entire height and width of the front 
end plays a particularly important role. On the A- 
class this purpose is served by the second impact 
plane with its stable transverse connections to the 
lower support structure. Even in a frontal collision 
with a larger car, the occupants of the smaller one are 
not at a major disadvantage in this case. 

The strengths of the sandwich concept of the 

Fig 7: 





car are not confined to front-end crashes. They also 
apply to side and rear impact collisions: in the side- 
on crash the other car hits the A-class at the precise 
level where the body of the A-class is at its strongest, 
at the level of the load-bearing structure for the floor. 
The occupants are seated above this impact level and 
are therefore optimally protected. The load bearing 
structure of straight cross members. Fig 7, and side 
members is better able to contain and reduce the 
impact energy than in a car of conventional body 
design in which the impact mostly occurs above the 
floor structure. 

The short crumple zones at the front of the A- 
class make it necessary to activate the belt tensioner 
immediately after the crash. As a result, the occu- 
pants not only take part earlier and longer in the 
deceleration of the vehicle, but they also have a 
longer distance for their crash-induced forward move- 
ment. The earlier the safety belt builds up its restrain- 
ing force, the longer the distance available to the belt 
and belt tensioner in which to arrest the forward 
movement of the occupants. In this way, the forces 
unleashed on the occupants are low, despite the 
massive deceleration of the occupant cell. This also 
serves to compensate for the design disadvantages of 
short deformation zones in small and compact cars. 
The innovative restraint system in the A-class in- 
cludes full-size airbags for driver and front passen- 
ger. The airbags fitted as standard have a volume of 
64 and 130 litres. They are operated by a new 
propellant technology which offers certain advan- 
tages when recycling: the gas generator for the driv- 
er’s airbag is filled with an acid-free solid propellant, 
and the passenger airbag contains an environmen- 
tally safe liquid gas mixture of helium and argon, 

The car has inertia-reel seat belts with belt 
tensioners on the front seats: newly developed com- 
pact belt tensioners compensate for seat belt slack 
and ensure that in a crash the occupants are even 
better tied to the passenger cell by snugly fitting 
belts. The belt strap is tensioned by small steel balls 
in the seat belt retractors which, when the belt 
tensioner is activated, are set in motion and rotate 
counter to the belt strap shaft. In this way they first 
stop the belt strap and then coil it back up at high 
speed. This drive system has two outstanding advan- 
tages: high tensioning capacity and compact con- 
struction of the seat belt retractor. 

The inertial reel seat belts for the front seat 

occupants are height-adjustable in three stages. The 
seat belt retractors of the front seat belts incorporate 
torsion bars which in a crash are slowly deformed 
from a certain force upwards, thereby reducing the 
locking action of the inertial reel seat belts. This 
reduces the risk of injury to the occupants who 
moreover, thanks to the force limiter, can sink very 
gently into the airbags. Belt force limiters and airbags 
are thus designed to function as part of a precisely co- 
ordinated system. The side airbags in the front doors 
form part of the standard specifications of the A-class. 
A newly developed gas generator with liquid gas 
filling is also used for these. M-B also fits belt 
tensioners as standard on the outer seats of the rear 
seat unit. These ensure that in a crash rear seat 
passengers also benefit in good time from the decel- 
eration of the body structure. The height adjustment 
of the inertial reel seat belts functions automatically 
on the rear seats. As on the C- and E-class cars, child 
seats which fold out of the rear seats at the touch of a 
button are available as an option. 

Primary safety system 

For the first-launched vehicle, the front axle has a 
modified McPherson suspension system with coil 
springs, twin-tube shock absorbers and torsion bar 
stabiliszer. Fig 8. The suspension components are 
mounted together with the rack-and-pinion steering 
gear, engine and gearbox on a frame-type integral 
support that is bolted to the body at eight points. 
Unlike the original McPherson suspension, the anti- 
roll bar of the A-class does not take part in wheel 
location but is linked to the suspension strut via 
plastic suspension: the results include better elasto- 
kinematics and increased rolling comfort. 

Ride comfort is also enhanced by the low- 
friction suspension struts and damper guide units, the 
Teflon coating of the front bearing of the torsion bar 
and the various noise-damping measures. All wheel 
location, suspension and damping components have 
been weight-optimized, such that engineers have 
achieved an exceptionally good balance between 
load-bearing ability and intrinsic weight: the effects 
of this balance only become evident in the sum of the 
car’s many detailed solutions. 

Another conceptual advantage that is felt in the 
handling of the car is the position of the steering gear 
in front of the centre of the wheel: this layout is only 



possible because of the sandwich double-layer floor 
concept, with the engine and gearbox located partly 
underneath the floor panel. Steering response is 
improved by this geometry. 

A compact servo pump for the steering of the 
A-class, for the first time in a Mercedes is not 
mechanically driven but is powered by an electric 
motor with electronic control. Pump capacity can be 
adjusted precisely to suit demand and the instantane- 
ous traffic situation. For example, when driving on 
the straight, the electric motor runs at reduced speed 
and consumes less energy. The three-spoke steering 
wheel of the car is a new development and is designed 
in particular to satisfy the needs of lightweight con- 
struction. The 370 millimetre diameter steering wheel 
is an aluminium pressure-casting, with a rim of 
polyurethane foam coated in plastic or leather, de- 
pending on the equipment level and customer option. 
The steering wheel material crumples in a controlled 
way in the event of high-force collision impact. 

The company opted for trailing arm suspension 
with coil springs, torsion bar stabilizer and single- 
tube gas-pressure shock absorbers. The main advan- 
tage is that trailing arm suspension can be designed 
with all its components underneath the load floor, so 
that it does not encroach on the interior space. The 

shock absorbers and springs are located in an other- 
wise unused space in front of the wheel centre. With 
this geometry, the suspension components support 
the straight side members in a crash, thereby also 
improving occupant safety. 

The dynamic qualities of the rear suspension 
are particularly due to special tie-bolts that control 
the elastic deformation of the side members and 
combine with them to form a rectangular linkage 
structure, Fig 9. Changes in the angle of the side 

Fig 8: Front suspension 

Fig 9: 





member and wheel about the vertical axis are com- 
pensated with up to 75 percent efficiency. The sus- 
pension components, side members, coil springs and 
shock absorbers are mounted on a suspension subframe 
in a single compact block that is fixed to the body- 
work on four sound-proofed rubber bearings, partly 
with hydraulic damping. The rear suspension unit is 
mounted with the aid of four vertical assembly bolts. 
None of the suspension components projects up- 
wards above the level of the floor. 

The technology and layout of the brakes, Fig 
10, correspond to the particular features of the A- 
class drive system and vehicle concept. Engineers 
opted for floating-calliper disc brakes at the front and 
drum brakes at the rear, because this combination 
provides optimal braking safety, stability and endur- 
ance for a front- wheel drive car, say M-B. A brake 
booster and the anti-lock braking system provide 
added safety. The drum brakes on the rear wheels are 
characterized by a lightweight construction: the sup- 
porting body for the wheel brake cylinder and brake 
shoes, the anchor plate and brake cylinder are made 
of aluminium. In order to satisfy the high safety 
requirements, the hubs are made of steel and the 
brake cylinders of grey cast iron. An automatic 

adjusting system compensates for brake pad and 
drum wear. The parking brake in the A-class is 
operated manually, using a lever between the front 
seats. The control cables are maintenance-free, since 
they have automatic length adjustment. 

ESP keeps the A-class directionally stable and 
reduces the risk of skidding when cornering by 
selectively applying the front and rear brakes. While 
driving, the ESP computer compares the actual be- 
haviour of the vehicle with the calculated set points. 
If the car is deviating from the safe ‘ideal line’, the 
system responds at high speed with a specially devel- 
oped logic, bringing the car back to the right track in 
two ways, firstly by the precisely dosed braking of 
one or more wheels and/or by controlling engine 
torque. By these methods, ESP corrects both driving 
errors and skidding movements caused by slippery or 
wet surfaces, loose chippings or other difficult road 
conditions that normally give the driver very little 
chance to keep the car in line by steering and braking 
action. The A-class is the first and only car in the sub- 
compact and compact car class to have such an 
innovative driving safety system. The Electronic 
Stability Programme also includes the compnay’s 
electronic Brake Assist System. 

Fig 10: Braking 



Ford Focus 

Ford’s Focus, Fig 11, brings the company’s 21st 
century hard-edge styling to its most mainstream 
model and also marks a completed market reposi- 
tioning of a class of vehicle starting as the Mark 1 
Escort to a vehicle refined enough to compete in the 
neo-luxury sector. The refinement of individual sys- 
tems to achieve the transformation is described be- 

Friction within the front suspension struts is 
reduced by off-set coil springs that eliminate bending 
forces acting upon the strut, Fig 12. Zero off-set 
geometry — via new A-arms supported by stiff, front 
A-arm bushes — provides precise lateral wheel 
control for enhanced precision and on-centre steering 
definition. Large, rear A-arm bushes permit longitu- 
dinal compliance for reduced impact harshness and 
improved ride performance. Negative scrub radius 
(-12 mm) enhances braking stability, particularly on 
surfaces in which available grip varies between the 
left- and right-hand tyres. Low steering friction lev- 
els minimize resistance and stickiness for improved 
feel, response and driver confidence. . Body-roll is 
controlled by anti-roll bars, which are 20 mm diam- 
eter front and rear. Large, pre-loaded bushes provide 
immediate linear control. Steering input induces 
immediate directional change (turn-in) with very 
little body roll. Ultra-stiff cross member, attached 
directly to floorpan, increases the precision of geo- 
metric wheel control and further stiffens the body. 
Friction within the steering system has been reduced 
20 per cent and damping friction reduced, Fig 14. 

Adoption of a fully independent. Control Blade 
multi-link rear suspension system, though similar to 
the system used on the Ford Mondeo estate, is rede- 
signed for Focus. Here the Control Blade makes 
extensive use of stampings for key components, 
which reduces cost, unsprung weight — by 3.5 kg per 
wheel — and assembly time, while offering even 
greater geometric precision, Fig 13. It has ride com- 
fort, steering precision, handling, braking and stabil- 
ity advantages, as well as NVH and package gains 
over the common twist-beam axle. There is greater 
longitudinal compliance without compromising lat- 
eral stiffness, which boosts steering precision, feel, 
handling and driver confidence. Intrinsic stability is 
provided by passive rear-wheel steer. The wider 
distribution and de-coupling of five separate load 
paths, compared with two for the twist beam design, 

Fig 11: Ford Focus 

helps to reduce transmission of noise, vibration and 
harshness levels to the passenger compartment. The 
control blade, spring link and toe link are all pressed- 
steel components. The camber link is the only cast- 
ing. Compliant bush provides recession, which al- 
lows rear wheels to move both upwards and rear- 
wards to absorb road bumps. 

An all new, highly rigid yet lightweight plat- 
form architecture (826 kNm/rad; 1070 kg for the 3- 
door — stiffest and lightest in class) provides best 
possible base for enhanced ride, handling, steering 
and active/passive safety. The combination is 
achieved through laser-welding, optimized panel 
gauges, extensive swaging, local reinforcement of 
critical points and optimized spot-weld distribution. 
Cross-member and powertrain pick-up points are as 
much as 150% stiffer than surrounding areas. The 
lightest model weighs just 1070 kg; both lighter and 
a full 15 per cent stiffer than all recent class entrants. 

The rear side rails, for example, are tradition- 
ally made from four separate parts. For Focus, these 
parts have been replaced by a single laser-welded 
blank, in which the gauge reduces progressively in 
three stages from front to rear. The result is a structure 
that is stiffer, with more rigid mounting points for the 
suspension, has improved crash performance and 
saves more than 1 kg per rail in weight. Similarly, the 
gauge of steel in the B-pillar is 2.25 mm thick at the 
top but tapers to just 1.1 mm at the base, providing a 
stiff upper section for a controlled and superior 
performance in the event of a side-impact crash. 

Zero offset geometry gives total separation of 
load paths, with lateral forces being fed through the 
stiff forward A-arm bush (1) while longitudinal forces 
are controlled by the more compliant rear A-arm 
bush (2) shown in Fig 1 3 . Large-diameter coil springs 
are offset, to feed suspension loads directly to them, 
eliminating bending forces on the struts for reduced 
friction and superior responses, Fig 14. 



Fig 12: Large diameter coil springs are offset so 
that suspension loads are fed to them without 

Fig 13: Zero-offset geometry allows lateral forces 
to be fed through bush (1) while longitudinal ones 

Stability is enhanced by maintaining optimum 
braking performance to the rear wheels in all condi- 
tions, whether the vehicle is heavily laden or carrying 
only the driver. On non-ABS-equipped vehicles, this 
is provided by a pressure cut-off proportioning valve. 
For the 1.8- and 2.0-litre petrol and diesel estate, a 
load-sensitive pressure limiting valve is used. 

When ABS is specified, brake proportioning is 
achieved using the electronic brake-force distribu- 
tion (EBD) system. It uses the ABS sensors to moni- 
tor the speed of front and rear wheels and adjusts the 
pressure distribution to help the rear wheels retain 
traction. EBD is self-compensating for all load con- 
ditions and imperceptible to most drivers. It is de- 
signed to reduce stopping distances and improves 
vehicle stability in conditions that fall short of trig- 
gering the ABS. ABS also forms the basis of the car’s 
Traction Control System (TCS) and Electronic Sta- 
bility Programme (ESP). TCS is a development of 
Ford’s dual-mode system with low-speed brake in- 
tervention coupled to a cut in engine power, above 3 1 
mph, delivered via the EEC-V engine management 

ESP acts whether or not the driver operates the 
brakes. In tests, it maintains vehicle stability at much 
higher speeds than would be possible without it — 
thereby placing less demand upon driver skill and 
reducing stress in extreme conditions. ESP uses 
software to control engine power and brake indi- 
vidual wheels to restore stability when the driver has 
exceeded or is beginning to exceed the car’s dynamic 
capabilities. Driver’s intentions — direction of travel, 
measured via a steering angle sensor — are compared 
against vehicle behaviour. Vehicle behaviour is moni- 
tored constantly by three additional sensors for yaw 
rate (rotation), lateral acceleration (centrifugal force) 
and wheel speed, with the latter sensors being inte- 
gral to the ABS. Yaw control is achieved via the 
vehicle’s ABS, with braking being directed to indi- 
vidual wheels through the ESP modulator. ESP uses 
information from three sensors and the four ABS 
wheel-speed sensors to determine the cornering be- 
haviour. It interactively compares the data in a dy- 
namic handling map stored in a special on-board 
computer. Additional ESP sensors monitor steering 
angle, lateral acceleration and yaw moment, which 
denotes the vehicle’s tendency to turn about its 
vertical axis. As soon as any tendency for the vehicle 
to move off the chosen line is detected, engine power 



is reduced, and individual brakes are applied momen- 
tarily to bring the car back on course. 

Powertrain With the Zetec SE low-friction 
engine, Ford used high-compression ratios, knock 
sensing and low idle speeds to deliver a 25 per cent 
real-world fuel economy improvement over past 
engines. The existing 1.4-litre Zetec SE unit, now 
fully re-calibrated for Focus to serve as a fuel- 
efficient 75 PS power unit, is supplemented by an all- 
new all-alloy 100 PS 1.6-litre Zetec SE engine. All 
engines require servicing only every 10,000 miles, 
with spark plug renewal every 40,000 miles and 
valve clearance attention only every 100,000 miles 
on petrol versions. All also feature a new torque roll 
axis (TRA) mounting system, which helps reduce 
powertrain shake on rough roads, results in a smoother 
idle quality and greatly reduces the degree of noise 
and vibration transmitted from the drivetrain into the 
bodyshell, Fig 18. Internally, the Zetec engines fea- 
ture an alloy ladder frame between the base of the 
block and the crankcase, Fig 19,which boosts rigidity 
of the powertrain assembly by 30 per cent. The alloy- 
block Zetec SE has a similar stiffening member 
added as a bearing beam within the crankcase skirt. 
These devices — together with new lightweight 
pistons and con rods — reduce second-order shaking 
forces by 20 per cent. Newly designed cam covers 
and low-noise ancillary drive systems generate sig- 
nificant improvements in running refinement and 
reduce perceived engine noise by 50 per cent. 

Cables transmit fore/aft and left/right gear le- 
ver movements to the selector mechanism, eliminat- 
ing a potential path for NVH to the interior, Fig 20. 
Cable shift also is smoother than a rod-operated 
system. Hydraulic clutch requires lighter pedal effort 
with less travel for a smoother take up. An updated 
version of the Ford B5 five-speed manual transmis- 
sion has external ribs along the transmission and 
clutch housing to increase drivetrain stiffness for 
reduced noise and vibration. Internally, first and 
second gears are equipped with double-cone 
synchronizers for smoother downshifts. Selection of 
reverse is also slicker. Durability is enhanced by the 
incorporation of strengthened final drive bearings 
and a new synthetic oil lubricant, which also eases 
gear selection when cold. New sealed-for-life bear- 
ings and fluid seals protect from dirt particle ingress 
for extended transmission life. 

Now available on Focus is a four-speed auto- 

Fig 18: Powertrain mass is supported by two 
mounts in line with the virtual axis of rotation, a 
separate arm resisting torque 



matic transmission which features an overdrive-top 
and lock-up torque converter. It has been designed 
specifically for use in front-wheel drive applications 
and is unusually light and compact. It is controlled by 
an electronic synchronous shift control (ESSC) mod- 
ule linked to the EEC-V engine management system. 
A centrally mounted quadrant selection lever has six 
positions (P,R,N,D,2,1), while overdrive can be 
switched by a separate, thumb-operated push-button 
at the side of the lever. The ESSC works in conjunc- 
tion with the EEC-V using information read from 18 
different engine and transmission sources to calcu- 
late the best possible shifting strategy for the driving 
and operating conditions. Fuel economy and per- 
formance are close to those of the equivalent manual 
transmission models, while the system’s shift quality 
and speed of response set new standards for this class 
of vehicle. 

Body structure and systems Focus meets more 
than 100 different real-world crash tests, modes and 
events, well beyond existing or currently proposed 
legislation, including all mandatory European and 
US destructive tests. At the front, beyond the bumper 
support beam, the car’s cross member is made from 
ultra-high-strength steel, so it is robust and stable and 
will not tear upon impact, Fig 21. Resistance to 
tearing helps spread crash energy into the car’ s lower 
side rails, which is particularly effective in offset 
impacts. Additional to the progressive collapse of the 
lower side rails, which absorbs most of the energy 
ahead of the passenger safety cell, a secondary load 
path channels the remaining energy over the roof and 
around the doors to ensure safe evacuation of passen- 

At the rear, the objective is to minimize intru- 
sion of the passenger compartment in a rear impact, 
while also ensuring that the doors remain operable 
and the integrity of the fuel system is maintained. 
This is achieved via the progressive collapse of the 
rear side rails, supported by the wheel housings. 
Stamped and laser-welded, the side rails are made up 
of three different metal gauges — changing in thick- 
ness from 1 .6 mm to 2.6 mm (in the area of the fuel 
tank) and 2.4 mm — to provide progressive control- 
led collapse. The compound structure avoids over- 
gauging near the rear bumper and eliminates the need 
to add reinforcement further back. Vehicle weight is 
thus reduced, by 1 kg per side, and the number of parts 
and welding operations are reduced, improving as- 
sembly tolerances and providing robust control of the 
collapse performance. For side-impact protection, 
the entire side structure, which includes door beams 
and anti-burst door latches, is designed to absorb 
energy and reduce intrusion, especially above the 
door bars. To achieve this, the B-pillar is made from 
a laser-welded, high-strength steel blank, with a 
gauge thickness that reduces from 2.25 nun at the top 
to 1 . 1 mm at its base. Additional inner reinforcement, 
designed to increase rigidity, makes it highly resist- 
ant to ‘kinking’, so it can more effectively channel 
crash energy into both the rocker and roof rail, and 
support the side airbag during deployment. 

All elements of the Focus restraint system 
work in harmony so that the whole structure is 
integrated and fine-tuned to provide maximum pro- 
tection. This starts with the locking of the inertia seat 
belt reels which occurs approximately 6 milliseconds 



after impact. (All timings depend upon the nature of 
the impact, whether it be straight, angled, offset, 
against a solid object or another vehicle). Approxi- 
mately 4 milliseconds later, before the head and chest 
begin to move forward, the airbag sensors trigger the 
front airbags and the small pyrotechnic charges in the 
pre-tensioners, which eliminate any slack in the seat 
belt and negate the effects of bulky or layered cloth- 
ing. The front airbags take approximately 35 milli- 
seconds to fully inflate, which is completed just 
before the occupant meets the bag, approximately 50 
milliseconds after impact. To reduce chest loading 
and minimize injury, Focus is equipped with load- 
limiting devices within the inertia reel assemblies. 
These pay-out belt webbing after a predetermined 
load is exceeded. This process is progressive and 
occurs between 45 and 70 milliseconds after impact, 
ensuring that the cross belt pressure against the 
occupant’s chest remains fairly constant. 

The load-limiting device consists of a torsion 
bar within the traditional inertia reel, and the typical 
pay-out is 150 mm. The system is calibrated to help 
ensure that front seat occupants do not meet the 
airbag until the optimum point, which is just as it is 
beginning to deflate. The progressive pay-out also 
ensures that the upper torso is less likely to twist, an 
action that can compromise the efficiency of the front 
airbag. Seatbelt pre-tensioners are designed to pull 
the buckles down and take up any slack, optimizing 
their position on the torso and maximizing the re- 
straint provided by the belt webbing. 

Focus’ new head-and-chest combination side 
airbags provide greatly enhanced head and chest 
protection in side impacts. In addition to acting as a 
barrier between the occupant and the side of the 
vehicle, the tall, 15.5-litre side airbag — stored in the 
seat back’s outer side bolster — is designed to 
cushion the head, minimiszing lateral head injuries. 

The car is equipped with the latest version of 
Ford’s Safeguard Passive Anti-Theft System (PATS). 
Safeguard is a Thatcham-approved high-security 
engine immobiliszer operated via a miniature trans- 
ponder within the key head. It transmits a unique 
coded signal, with several trillion possible codes, to 
a transceiver situated around the steering lock. Once 
the key is identified as correct, the engine is mobi- 
lized. Without it, the engine will not start and is 
effectively dead. PATS offers several advantages 
over competitive systems. The system arms itself. 

Fig 22: Seats are designed to attenuate vibration 
in the 10-50 Hz frequency range 



eliminating human error through forgetfulness. The 
interception of the key code by a thief is virtually 
impossible because the transponder only has the 
power to pass the code by radio to the transceiver 
when it is within the electrical field of the transceiver. 
PATS also incorporates a special erasing procedure 
and smart ‘learning’ mode to maintain security after 
resale and eliminate the requirement for a master key . 
In addition, a new, key-operated hood-release mecha- 
nism makes a potential theft even more difficult. By 
eliminating access to the engine compartment via the 
traditional cabin lever and cable, to ensuree that a 
would-be thief could gain access to the engine or 
alarm siren only with the car’s ignition key. 

Fig 21: High specific-stiffness body shell 

Fig 23: Large diameter rear A-arm bushes give 
longitudinal compliance without compromising 
lateral stiffness 

The car’s shrouded lock mechanisms and wir- 
ing present the first obstacle to forced entry, render- 
ing break-in many times more difficult. A further 
obstacle is posed by a double-locking door system, 
which is linked to the car’s central locking and 
optional remote-control facility. Once the double 
locks have been set, only the key or remote button can 
release the locks. A further deterrent is presented by 
a tamper-proof perimeter and interior scanning alarm. 
Once armed, it detects the opening of any doors, 
tailgate or hood, as well as any movement within the 
passenger compartment. The remote-control locking 
system has a 10-metre range and uses a miniature 
transmitter working on a rolling frequency system to 
eliminate possible interception by an electronic scan- 

The body structure and seating were tuned to 
reduce resonant peaks and move natural vibration 
frequencies outside the operating range, Fig 22. Deep 
swaging of body panels, especially the floorpan, help 
to maintain stiffness while permitting down-gauging 
for weight reduction. Where this was not possible, 
around the bulkhead for example, dual-function dead- 
ening pads were developed, with a viscous layer to 
dampen sound and a mass layer to absorb it. As well 
as tuned suspension mounts, Fig 23, torque roll axis 
engine mounts also are used to reduce powertrain 
noise, improve idle quality and reduce powertrain 
shake on rough roads. Such a mounting system 
separates the powertrain mass from its torque reac- 
tions by using two rubber-to-metal bonded mounts 
located at either end of the assembly, in line with the 
axis of rotational inertia, plus a dedicated torque 
reaction link. The front mount incorporates a hydrau- 
lic chamber so the modulus of the rubber can be 
optimized to reduce NVH transmission at high en- 
gine revs, without reducing idling smoothness. 

Two direct acoustic paths from the drivetrain to 
the interior have been eliminated with the adoption of 
a hydraulic clutch mechanism and cable-operated 
gearshift. A unique flexible exhaust joint, situated 
behind the engine manifold, limits the transmission 
of drivetrain vibrations along the length of the ex- 
haust system. It also has specially tuned exhaust 
hangers positioned at the natural vibrational nodes to 
provide further isolation. ‘Helmholtz’ resonators are 
used to filter or select individual frequencies so that 
an appropriately purposeful, yet subdued sound is 



Land Rover Freelander 

With prices pitched at under £16 000 the Freelander, 
Fig 24, has finally brought Land Rover within grasp 
of the profitable small sports-utility-vehicle market 
and first results are showing the vehicle to be per- 
forming well financially. Trade sources have pro- 
jected higher residual values than the Honda CRV 
and Toyota RAV4. It is the first of the company’s 
vehicles with a unitary construction body shell and 
transverse power unit and a well-planned engine 
compartment gives good service access. Fig 25. 


The integrated underframe comprises high-strength 
steel box-section longitudinals, with eight 
crossmembers, all welded through to the floor pan. 
Box-sectioned outriggers from the crossmembers 
connect the side sills through to the centre tunnel 
member. Front and rear suspension sub-frames are 
bolted rigidly to the shell to enhance its torsional 
rigidity. The upper shell involves monoside pressings; 
there is a deep rear-heelboard serving as a further 
crossmember; there is a strong integral fascia panel, 
too; all these combining to enhance the stiffness of 
the superstructure. For the three-door variant the 
torsional figure is 13,500 Nm per degree and rises to 
17,500 on the Station Wagon. 

Within each lower A-post there are five dia- 
phragm pressings to stabilize the box-beam and 

provide additional crush resistance under impact. 
Front-ends of the underframe longitudinals crush 
progressively on impact, after initial deflection of 
bumper armature and crush cans, a tapered box- 
section in the top of each engine bay side-panel also 

Fig 24: Land Rover Freelander 



absorbing impact. In more severe crashes, bending of 
the curved portions of the main longitudinals takes 
place. The suspension towers are braced back to the 
strong A-pillar area via box-sectioned members ei- 
ther side of the bulkhead. The rear bush mountings 
for the lower front suspension arms are designed to 
break free of the subframe, under severe impact, to 
minimize intrusions into the front footwells. From 
the bulkhead backwards the floor structure has a 
‘branching’ layout to help spread impact forces as 
widely as possible, with a curved ‘horn’ box-member 
sweeping back from the longitudinal to the central 
tunnel area, the box-sectioned side sills also being 
brought into play. 

Body systems 

A spoiler-type electrically operated glass sun-roof 
has been engineered to tilt open at the rear and slide 
backwards above the roof panel, to maximize rear 
headroom, two-stage slides permitting 70% clear 
aperture opening against the usual 50%. At the front 

of the vehicle’s roof there is a twin-panel targa 
structure, with detachable T-bar. Panels can be indi- 
vidually tilted open at the trailing edge, or taken out 
— at which point wind deflectors spring into posi- 
tion. Detached panels stow in a bag on the back of the 
rear seat squab. The complete targa unit is con- 
structed as a cassette, having a one-piece SMC mould- 
ing which is bonded to the bodyshell with a poly- 
urethane seal. Together with a number of the interior 
mouldings this is made by a gas injection moulding 
technique to increase specific strength and stiffness. 

Roof modules for the 3-door vehicle come in 
either softback or hardback forms, both of which 
involve the same tail-door electrically operated drop- 
glass as the station wagon, Fig 26. Roofs are thus 
interchangeable for those customers wishing to in- 
vest in the alternative. On the softback variant, Fig 
27, the vinyl roof folds upwards and furls within an 
integral tonneau at roof level so rear vision is unim- 
paired. PVC sidescreens are zipped in place and can 
be individually removed to leave the rear roof as a 

Fig 26: Station-wagon variant 



canopy with strong air through-flow. Key body di- 
mensions are tabulated in Fig 27. 

A green-tinted laminated windscreen is bonded 
in place and all other window glass is Pilkington 
Optikool solar-control glazing, cutting solar trans- 
mission by 50%. Front wing panels are in an engi- 
neering polymer chosen for its damage resistance and 
1.5 kg per wing weight saving. They are jointly 
painted with the main shell and have to withstand 1 80 
C stoving temperature. Headlamp lens too are poly- 
mer-mouldings, a silicon-coated polycarbonate be- 
ing used to allow a complex optical pattern which 
maximizes beam control. The front bumper and 
grille surround is a large polypropylene moulding 
mounted on an aluminium armature while that at the 
rear is mounted on a compression moulded glass-mat 
reinforced polypropylene member. Also in 
polypropylene are wheelarch mouldings, wheelarch 
liners, lower sill flange finishers, tail-door lower 
edge finishers and mudflaps. 

Interesting interior detail is the facia-mounted 
rocker switch construction. Fig 29. The switch con- 
tact metal is integrated during the injection moulding 
and a common base moulding and contact plate are 
used. Switch warning lights are LEDs rather than 
bulbs. Another detail is in the driver seating, which 
has an adjustable lumbar support that does not stiffen 
the squab springing when adjusted outwards. 

The side-hinged tail-door carrying the spare 
wheel also involves a specialized construction tech- 
nique, for mounting the wheel without bridging inner 
and outer skins over the space required for the drop- 
glass. A large die-cast aluminium drum-moulding 
carries the wheel and spreads the mounting bolt loads 
over a very wide area. A neat support arm for the 
high-level stop-lamp is integrated into the drum. A 
gas-spring strut beneath the door has ‘keeps’ at half 
and fully open positions and facilitates operation by 
users of all sizes. The drop-glass has a grooved seal 
at its top edge and electrical interlocks are provided 
alongside a short-drop/lift control system to release 
the glass from the groove when the door is opened and 

Vehicle mechanical systems 

A key quality of the vehicle is the provision of good 
off-road performance while maintaining on-road han- 
dling comparable with the best of conventional car 

practice. Rover Cars experience was drawn upon for 
the independent suspension system involving 
McPherson struts on all four wheel stations which 
provide vertical travels of 7 in front and 8 in rear. 
There are high levels of fore-aft compliance, around 
twice the normal 5 mm. Steer fight is avoided by 
carefully designed suspension bush configuration, 
aided by a steering rack with centre take-off, giving 
very long track rods. An anti-kickback valve in the 
power steering hydraulics eases the shocks of cross- 
country driving, Fig 30. While disc brakes are used at 
the front, the rear wheels rely on drum brakes to 
provide park-brake performance appropriate to the 
extremes of cross-country operation. The ABS has a 
special control system which builds in electronic 
traction control (ETC) and hill-descent control (HDC) . 


Overall lenylli 

1382mm (including 1 5" spare wheel) 

Overall width 

{iiidi.i(iifiy iriinuis) 

Overall width 

1805mm (wtclucmg mirrors) 

Ovetull li ci gill 

1 F!i/nim (Including root tails) 

Wheel IMSv 


Truck front 

t via null 

Track rear 


Miliinuni ptound clearance 

iswnm (to trpril 'axle') 

ADprwctvweMepai! Mule 

EC 3C*/2'IW 

From overhang 


Near overhang 

OOlntn (insludmg 16" spare wheel) 

Maximum from leg room 

1 PR/lir 

Minin ini rear leg wum 

93d inm 

f mill head rnurri 


Hear head room 

101 Omni 

Front shoulder room 

1 4PCmm 

Hear shoulder rnnni 


l (Kidspace height 


Loadspecn length sente up 

44" (mil 

1 MdstHcts Liiglh scats folded 


Itndspace width 


Luadspaee width between arches 


Loadspa.-.a sill height 


l oadspace to tool wilt seats up 

fititl lilies 

Loadspace In mot seats lokled 

1313 Hires 

Luadspaee lu glass - seats up 

3/1 litres 

loadspar? to gloss scats folded 

722 litres 

L'jadspucc door upcrlum width min 


boadspace door aperture widlli mux 



l litres 

Fuel lank capacity 

13 ualluies / 09 litres 

Fig 27: Key dimensions 



Fig 28: Softback variant 

ETC adds a further degree of limited slip control, to 
that provided between front and rear axles by the 
viscous coupling, but across each axle. HDC pre- 
vents tobogganing on sharp downhill brake applica- 
tions by a hands/feet-off approach in which engine 
overrun at idle is detected and brakes gently applied 
to maintain maximum descent speed of 4.4 mph. 

The powertrain involves either the K- or L- 
series engines, the K being specially tuned to provide 
power peak 250 ipm further down the speed range. 
The engine is also the first to use a retumless petrol 
injection supply system, with more sophisticated 
electronic control over the in-tank fuel pump. The L- 
series diesel has water-cooled EGR to reduce NOx 

Fig 29: Interior features 

Fig 30: Steering system 



Project Thrust SSC 

G. T. Bowsher, Chief Designer (Mechanical) Thrust 
SSC Project, was responsible for spaceframe/struc- 
ture/wheels/brakes/suspension/steering systems/en- 
gine mounts/ throttle systems. Seen in Fig 3 1 with the 
Dunlop-made wheel and brake, he described the 
engineering to a contemporary specialist journal 1 . 

He explained that Thrust SSC, although en- 
tirely new in concept, with twin jet engines and rear 
wheel steering , incorporates technology developed 
in Thrust 2, which regained the world land speed 
record in 1973. A major feature of the rolling gear is 
the adoption of solid aluminium alloy wheels and a 
power hydraulic braking system which uses two disc 
brake callipers acting on each specifically designed 
high speed carbon disc. The Thrust 2 experience gave 
valuable information on the safe ‘footprint’ size for 
transferring vehicle weight to a desert surface. It also 
made it possible to predict how wheel brakes might 
effectively be used to stop the vehicle. This is be- 
cause the same rolling interface applies (aluminium 
alloy wheels and the low adhesion, friable surface of 
Black Rock desert where the very high speed runs 
were carried out). The shape and ground clearance of 
the new car dictated a wheel diameter of 34 inches 
(0.86 m). At the anticipated peak design speed of 
8400 rpm this means that the radial acceleration at the 
wheel periphery is 34,000 g — effectively making 1 
gm of material weigh 34 kg. 

Dunlop undertook the manufacture of all 
wheels for the jet car, both the high speed solid alloy 
sets (with spares) and the low speed wheels which 
utilized the main undercarriage tyres of the Lightning 

Fig 31: Chief Designer G.T. Bowsher with Dunlop 
wheel and brake 

fighter. This proved a major contribution at an early 
stage, and at a time when credibility for the project 
was of great importance. To produce the high speed 
wheels solid L.77 forged aluminium blanks supplied 
by HDA Forging were machined to provide a rolling 
tread width of 10 inches (0.25 metre) for the front 
wheels and 6 inches (0.152 metre) for the rear, 
steering, wheels. This is in line with the weight 
distribution of the vehicle’s 10 ton static weight. 

A straight web design for both front and rear 
wheels became essential in order to minimize stress, 
but for the front wheels, with the wider tread profile 
and larger peripheral mass, it also became necessary 
to omit the traditional central hole and have a solid 
centred wheel; only then could the high speed stresses 
be kept within the limitations of the L.77 material. In 
terms of the wheel installation this did not matter, as 
the decision, already taken, to steer Thrust SSC by the 

Figs 31/32: Front and rear wheels after final machining 



rear wheels and not the front meant that the central 
hole was not essential: its principal use would have 
been as a location for machining. As the rear wheels 
were to be steered a central bearing/stub axle hole 
was a necessity, but this could still be accommodated 
within the material’s stress limits because of the 
narrower tread width and lower peripheral mass. 
Three dimensional computational finite element stress 
analysis techniques were used in the design of the 
wheels, minimizing errors before a commitment to 
costly machining. Of interest, the rear wheels are 
slightly higher stressed than the fronts at any given 
speed, indicating the effect of the central hole. 

The finalized wheel shapes are given in Fig 3 1 
for the front wheels and Fig 32 for the rear wheels. Fig 
33 shows the three dimensional stress contours of a 
front wheel at 8000 rpm., the darkest shade indicating 
the highest stresses (note those in the bolt holes) and 
the lightest the lowest stresses, with the other shades 
giving intermediate values. All wheels should offer 
minimal resistance to rolling and yet have a high 
lateral resistance in order to provide good vehicle 
control. Three circumferential grooves on the radiused 
profile of each front wheel tread provided for this, 
allowing a progressive take up of wheel load and for 
desert surface deformation to give the required lat- 
eral stability. Twin keels were a feature of each rear 
wheel, with the tread shape again profiled to give a 

progressive take up of load, these were intended to 
penetrate the desert surface and allow the generation 
of lateral steering forces, when steering took place, in 
addition to providing good lateral stability to the rear 
of the jet car. The wide spacing of the keels improved 
the damping capacity at the desert surface interface 
and helped prevent wheel ‘shimmy’, a feature of the 
first set of rear wheels, which exhibited only a single 
central keel of low damping capacity. 

Although a parachute was the primary means 
of stopping the jet car from high speed. Fig 34,wheel 
brakes were essential to cover the lower speed range 
(200 mph to rest), when the parachute loses much of 
its effectiveness. In an emergency (if the parachute 
failed), the wheel brakes would need to stop the 
vehicle from over 300 mph. 

In most respects, the wheel brakes for Thrust 
SSC were laid down as a design fourteen years before 
following the conclusion of the Thrust 2 project: the 
best features of that design were retained whilst other 
features were design-developed as a natural follow 
on and improvement. It was simply a question of 
waiting for a new vehicle. Each wheel brake com- 
prises a wheel driven disc/rotor and a pair (one 
leading and one trailing) of brake callipers which 
hydraulically clamp the disc between suitable fric- 
tion stators (brake pads). The brake callipers react the 
friction drag loads developed between the brake pads 



and disc and pass them into the vehicle’s structure, 
whilst the brake disc absorbs the vehicles kinetic 
energy in the form of heat. The twin callipers equal- 
ize the brake loads in the system, particularly those 
reacting through the wheel bearings, and much im- 
prove the heat distribution in the brake disc. 

A critical element in the brake design, the 
author explains, is the drive junction between the 
wheel and brake disc. That of Thrust SSC took the 
form of its predecessor’s brake in having drive slots 
of a particular shape machined into the inner periph- 
ery of the disc, and engaging these with drive lugs 
machined into the wheel carriers. The slots, of almost 
‘keyhole’ shape, allowed the disc to expand due 
either to high temperature or high rotational speed by 
sliding radially on the drive lugs whilst retaining its 
concentricity with the wheel, the precision of the slot/ 
lug fit ensuring this. The shape of the ‘keyhole’ drive 
slot minimizes disc stresses, either thermal or rota- 
tional, whilst allowing a failure mode which retained 
the integrity of the disc as a major safety feature. 

A degree of lateral compliance between the 
hydraulically actuated brake pads and disc is essen- 
tial, partly to provide for evenness in the lateral force 
clamp, but also to allow the brake pads to move clear 
of the disc when not in use. In the Thrust SSC’s 
predecessor’s brake this compliance came by later- 
ally springing the brake disc into a fixed position, and 
applying a pair of twin opposed piston callipers: the 
opposed pistons would move the brake pads into and 
out of contact with the brake disc, whilst the latter 
could still expand and contract radially as required. 
The webs of the Thrust 2 wheels had been cranked 
around the brake callipers, an arrangement com- 

pletely unacceptable for the straight webbed wheels 
of Thrust SSC. For the new vehicles brake callipers, 
the body is fixed, and the disc allowed to move 
axially on the slot/lugs in order to provide the neces- 
sary lateral compliance in connection with single 
sided hydraulic pistons. Now, the hydraulic pistons 
would move the outboard brake pad into contact with 
the disc, then move the disc laterally into contact with 
the fixed inboard pad in order to provide the brake 
clamp. This arrangement minimized the space re- 
quired for the fixed arm/fist element of the calliper, 
and fitted in well with the straight webs of the new 
wheels. In a laterally moving disc it is essential that 
all lateral forces are equal and equally disposed about 
the brake disc’s axial centreline, this ensures that the 
disc does not jam in the drive lugs when actuated. The 

Fig 35: Calliper and disk assembly 

AKSTS 5.3 
OCT 1 1196 
SUB -1 

CKX -1.172 
SMN -5.677 

SKX -259.271 
■I 5.677 
■ 33.654 

m 62.031 
I 90.208 
H 110.385 
| 146.562 

rr 1 174.74 

m 202.917 
■i 231.094 

Fig 33: Results of FE stress analysis 
on front wheel at 8000 rpm 



two callipers (one leading and one trailing), of the 
SSC brake arrangement provide for this, but each 
calliper has two pistons the pistons: are therefore 
diagonally connected on separate hydraulic lines, so 
that should an hydraulic line failure occur, the brake 
disc will still be evenly loaded in both the axial and 
rotational planes. The arrangement of the SSC brake 
is illustrated in Figs 35 and 36. Brake design was 
common throughout the jet car, with two brakes 
applied to each front wheel, and one for each rear 
wheel: brake effort therefore matched the dynamic 
weight distribution of the vehicle when decelerating 
on a low adhesion desert. 

As both the wheels and brake discs were in- 
tended to operate at much higher speeds than in the 
former jet car, the compacted graphite iron of the 
original discs could not be used again. Several alter- 
natives were considered, one being to minimize high 
speed stresses by reducing the mass density of the 
disc material. Of three possibilities in this area, one 
was the use of carbon/carbon fibre, used so effec- 
tively in aviation and race car brakes. It was with this 
in mind that Dunlop were approached in order to 
assess both the suitability of carbon as a disc material. 

and their interest in supplying such discs. 

In order to use this material, the brake pads 
would need to be from the same material, even 
though its thermal properties would ensure very high 
temperatures throughout the pad depth when com- 
pared with conventional materials. In view of the 
anticipated high brake pad temperatures two changes 
were made to the brake calliper design: the fixed 
calliper bridge/fist, originally specified in light alloy, 
became re-specified in steel, and the automotive 
style hydraulic pistons were changed to the Dunlop 
Aviation standard screw-in units, which were already 
fitted with a high temperature heat insulator. In 
addition, the brake discs were slightly reduced in 
outside diameter in order to use an available carbon 
disc blank, and the quantity of disc drive slots in- 
creased to sixteen. 

All rotary components from the wheel/brake 
assembly needed proof testing before being passed 
for use on the jet car. Quite apart from any theoretical 
analysis of wheel/disc shapes, spin testing to at least 
9000 rpm would ensure the use of the equipment at up 
to the 8400 rpm design speed. Spin testing took place 
at Pystock/Famborough, under the good offices of 
the DRA, following balancing by Schenck, and all 
components passed without problem, particular em- 
phasis on the post run inspections being directed to 
the relatively highly stressed wheel bolt holes. Be- 
cause of the relatively large inertia of the wheels 
compared with the available motor power of the 
spinning machine, a spin test from rest to 9000 rpm 
and back to rest took 35/45 minutes, with a full 5 
minute stress soak in the critical region of 8500-9000 
rpm. An additional test, for one brake disc only, was 
to spin test to destruction: this occurred at 10,500 
r.p.m.. For physical reasons, it was not possible to 
spin test a wheel to destruction. 

In operation the brakes, because of large inter- 
face friction variation with generated temperature, 
required a learning curve in order to get the best out 
of them. This done, the brakes have been trouble free 
for the whole of the project, and pad/disc wear has 
been small enough for only one complete set of brake 
parts to be used for all 66 runs completed. Although 
small, a disc/pad damage differential became pro- 
gressively noticeable between the front and rear 
brakes as the run numbers increased, with the rears 
taking the greater punishment, the likely cause was 
the heavily dust laden air, disturbed from the desert 



dust by the front wheels, passing with a scrubbing 
action between the rear pads and discs. Generally, 
brake temperatures peaked in the region of 450/500 
C, but in a parachute failed situation temperatures of 
800 C were reached, these temperatures were easily 
coped with. The high speed wheels were never oper- 
ated to their true operational limits, and therefore 
never achieved the maximum stresses anticipated 
from both rotational speed and wheel loading. Tread 
damage from stone impacts during the second Jorda- 
nian test series became a potential problem as speed 
rose to the mid-500’s: these damaged areaswould be 
stress raisers on an increasingly stressed surface at 
even higher speeds. 

Wheel interaction with the desert surface be- 
came a dominant feature of the Black Rock operation 
as speeds rose above 600 mph, and good vehicle 
control became even more critical as the variations in 
desert hardness and surface texture over a 1 3 mile run 
length played their hand, particularly as an interme- 
diate rainy period had changed the desert character- 
istics. In general, speeds up to 500 mph allowed the 
desert to imprint the shapes of the wheel treads, but 
through 600+ mph the desert beneath the wheels 
became very broken in their passing, possibly due to 
the formation of ‘impact’ shock waves. In the region 
of 650-750 mph the desert both beneath and either 
side of the car broke up considerably, probably due to 
the formation of large vehicle generated shock waves, 
those which were not airborne being dissipated by the 
desert itself. Beneath the wheels, to a depth which 
varied from 0.75 inch to 4 inches, the desert became 
pulverized into a fine dust: it was almost as if at high 
speed the jet car ran on a ‘fluidized bed’ rather than 
a solid desert, though sensors still indicated the 
dynamic variations in wheel load. In terms of desert 
firmness, miles 7-13 were best whilst miles 0-6 were 
softer with the measured mile being a combination of 
both: it was quite possible to feel the difference in 
desert drag over this length when driving a conven- 
tional vehicle. 

Intermediate rain before the final supersonic 
runs changed these characteristics, with reducing 
hardness, and with a particular detrimental effect on 
miles 4-6, where the surface became particularly 
friable. A total of 66 runs completed the programme 
for the Thrust SSC jet car, from the first tentative 40 
mph on the Famborough runway to the final pair 
which created the world’s first Supersonic Land 

Speed Record at 763.035 mph, (Mach 1.0175), and 
taking an intermediate LSR of 717.144 mph on the 
way. All runs were officially timed by the United 
States Auto Club under FIA rules. In the configura- 
tion used, with the Spey 202 engines the jet car would 
go no faster, as it could not penetrate the rapidly 
increasing supersonic drag. An engine change to the 
more powerful Spey 205 engines might have helped 
an increase in speed to 800 mph, but the true super- 
sonic objective had been reached, and it was decided 
to leave that mark to those who follow. 

Fig 36b: Rear brake 

Chapter 6: 

Systems development: powertrain/chassis 

A decade of patents specifications and company product 
releases provides an insight into automotive product 
development up to the period of new designs covered in the 
first five chapters. In the area of powertrain and chassis systems 
the trends to reducing both power losses and environmental 
pollution are clear in the examples which follow. 



Engine developments 

Performance ‘on spec’ 

Cosworth Engineering’s MBA high performance 
engine, first built as a technology demonstrator, is a 
high power-density unit which develops near maxi- 
mum torque over a very wide speed range and 
incorporates many race-leamt techniques. 

Having a dry weight of only 120 kg (overall 
dimensions 536 mm high X 20 wide X 490 long), the 
peak power of the MBA at 162 kW at 7000 rpm is 
commendable and the maximum torque of 256 Nm is 
developed at 4500; good in-gear acceleration can be 
obtained from 4000 rpm, in fact. Being produced 
independently of any client programme, consider- 
able detail is now available on its design and con- 
struction. The aluminium alloy unit is configured as 
a 2.5 litre, 24 valve, 90 degree V-6 and is notable for 
such innovations as inlet port throttling and an 
unusual head and block structure. The block itself 
weighs only 26 kg with its 87 mm bore and 70 mm 
stroke layout. This 1.24 B/S ratio allows the fitment 
of large diameter valves for high specific perform- 
ance but a lower ratio would be used for more 
emission-conscious variants. The block has a contra- 
rotating balance-shaft built into it which, running at 
engine speed, balances the primary couple inherent 
with this cylinder configuration. 

This shaft also has a centrifugal oil-separa- 
tor incorporated with it, Fig 1, as part of the advance 
crankshaft ventilation system. Oil is jetted to the 
underside of the pistons for cooling at high engine 
speeds. The head has pent-roof combustion cham- 
bers, giving 11:1 compression ratio, but the head-to- 
block joint is unique and involves angled head bolts. 

Fig 1: Vertical section through block showing 
balance shaft axis above that of crankshaft 

Camshafts are pre-assembled into separate housings 
which simplify the main head casting. Valves are 35 
mm diameter for the inlet and 30 for the exhaust and 
have an included valve angle of 40 degrees. The 
barrel apertures which form the port-throttles are also 
incorporated into the head. 

Cylinder block side walls are held in tension 
by the angled head bolts while the cylinder wall is in 
compression by reaction to the bolt load. Solid fire 
rings seal each bore while O-rings seal the water and 
oil galleries. Cross-bolting of the main-bearing caps, 
Fig 2, helps to contain costs by eliminating the need 
for tight-tolerance machining. Mating horizontal 
surfaces on the block and iron caps are first machined 
flat then the bolt holes are drilled and tapped. After 
bolting down the caps, dowel holes are drilled in the 
two vertical clearance slots between each cap and the 
block. Cross-drilled dowel pins, and then bolts, are 
then fitted prior to line boring the assembly. A two- 

Fig 3: Camshaft drive 



stage camshaft drive is featured with helical gear on residual exhaust gas to flow back into the inlet port, 
the cam nose driving two primary gears which inde- the resulting charge reduction lowering NOx emis- 
pendently chain-drive one pair of camshafts. The sion. An external EGR system would accentuate this 
two-stage reduction thus allows smaller camshaft process with external pipes. The MBA has a patented 
sprockets than usual, which reduces the overall size internal EGR system which circulates exhaust gases 

of the engine. Chain life is also extended with this rich in HC; this claims to reduce both NOx and HC 

slower moving layout, made quiet by individual emissions. A high performance engine usually re- 
cases moulded in special grade of nylon, Fig 3. quires long cam durations and a large valve overlap 

The crankshaft has offset pins to provide period and the resulting exhaust dilution is greater 

even firing, main and big ends being 50 and 42 mm than on a conventional unit. On the MBA the result- 

respectively. Connecting rods weigh only 0.49 kg ing rough running that this would cause is obviated 

and are 128 mm between centres; piston weight, too, by port-throttling which limits the reverse flows. The 

is kept down to 0.4 kg. The induction system com- depression created in the inlet port, during induction, 

prises a variable-geometry plenum which helps to recovers after the inlet valve closes, according to 

spread WOT torque over a wide speed band. Two speed and load conditions. As the inlet reopens there 

chambers into the plenum individually feed one bank is little pressure differential between inlet and ex- 

of cylinders; they are linked together by two pipes haust ports. The independent throttle mechanism, 
with switchable valves controlled by the engine Fig 4, on the MBA involves a port-throttle upstream 
management system. of the inlet valves, the tracts being fed into a common 

In a conventional plenum-throttled engine plenum. Thus, at a given speed and load, degree of 

there is pressure differential at part throttle between exhaust gas dilution is a function of inlet port depres- 

inlet and exhaust ports. Conventionally there is sion. Internal EGR is therefore controlled by the 

simultaneous opening of inlet and exhaust valves relationship between the two throttles, Fig 5. 

towards the end of the exhaust stroke which allows 

Fig 5: Cylinder head details (above left and right) Fi 8 4: Plenum and port throttles 



Advanced car and truck engines 

Lotus Engineering’s Jewel Project, Fig 6, seeks to 
meet some of the challenges of future engine require- 
ments. It involves a monobloc design to improve heat 
flow from the head to the main water jacket, the 
cylinder head having hardened surfaces interfacing 
with the aluminium alloy monobloc. Resulting re- 
duced temperatures for pistons and liners permit a 
safe raising of the compression ratio for greater 
engine efficiency. 

A hollow crankshaft saves weight and lowers 
inertial forces — also helped by the use of metal 
matrix composite conn-rods. Corrosion of such parts 
as water pump and thermostat housings is avoided by 
the use of GRP composites; fuel rails, too, are injec- 
tion-moulded for both lighter weight and better ther- 
mal properties for resisting fuel vaporization. Cam 
covers and outer walls of the coolant jacket are also 
of composite construction. A rubber joint isolates the 
monobloc casting from manifold and sump to reduce 
noise transmission. 

Scania has become the first truck manufacturer 
to undertake series production of a turbo-compound 
diesel engine, Fig 7, with an efficiency of 46%, over 
10% higher than the best gasoline engine figure. 
Exhaust energy is recovered in two stages; the first 
turbocharger (1) drives a compressor to force more 
air into the cylinders while the second is a power 
turbine (2) supplying additional power to the fly- 
wheel (4) through a hydraulic coupling and gear train 
(3). Power and torque are about 5% higher at 400 bhp 
and 1750 Nm. There is also a 6 gm/kWh saving in 
specific fuel consumption. 

At the left of the figure the power turbine is 
seen downstream of the turbocharger turbine, the 
latter being small in relation to mass flow at high 

~ . . TECO'GARY »^3ISE 

Project engine system 

engine speed — where, pressure ratio across the 
power turbine being high, it still provides enough 
output to drive the compressor. At low speed when 
pressure ratio across it is reduced, the compressor’s 
pressure ratio is maintained — partly because of the 
small turbine housing. Increased charge pressure at 
low speed thus makes it possible to inject more fuel 
and increase torque without air/fuel ratio becoming 
too low. Power turbine is connected to the crankshaft 
via a two stage spur gear set and hydraulic coupling. 
Mass inertia of the power turbine, converted to 
crankshaft speed, is about half that of a normal 
flywheel (gearratio 1:30); the ‘soft’ coupling reduces 
the peak torque amplitudes. The set-up is believed by 
Scania engineers to be a more satisfactory route than 
a variable-geometry turbocharger, some of which 
contain more than 100 moving parts and are liable to 
reliability problems. However, it is pointed out that 
greatest advantage comes with engines operating 
under heavy and sustained loads. 



Improved engine temperature 
management and differential 
expansion control 

During the cooling cycle of an engine, after combus- 
tion, there can be an unduly long retention of heat by 
some of the components. This heat can flow back into 
the incoming air-fuel charge with resultant loss in 
engine efficiency. The design in Fig 8 thus aims to 
make combustion chamber and charge induction 
surfaces store and release heat selectively, to limit the 
piston surface temperature and control thermal ex- 
pansion. The use of sprayed insulant coatings on the 
inside walls of engine combustion chambers is pro- 
posed here to manage the heat release from the 
combustion process. A low thermal expansion/high 
conductivity insert is also cast into the piston crown 
to inhibit differential expansion. An abradeable coat- 
ing is also sprayed on the piston walls, above the top 
ring, in order to reduce clearance and therefore 
enhance heat transfer. Under crown oil-spray cooling 
is employed, too, together with a dual wall construc- 
tion for inlet and exhaust manifolds, the net effect of 
which is to reduce the onset of knock and therefore 
allow the use of higher compression ratios to provide 
enhanced efficiency. 

The problem of differential thermal expansion 
of engine cylinder head and block, in different mate- 
rials, is said to be overcome in the design of Fig 9. An 
aluminium alloy head is secured to a cast-iron cylin- 
der block by extra-long bolts which engage with 
threaded bosses deep down in the block. Above the 
threaded portions the bolts pass through oil galleries 
and are warmed sufficiently to cause the correct 
linear expansion to compensate for that of the cylin- 
der head, between its support faces. 

Improved turbocharger 
waste-gate control and move to 
constant-pressure cycle 

A simple and reliable turbocharger pressure-control- 
ler, for providing enhanced operation of the waste- 
gate valve, is provided in the design of Fig 10 which 
permits standard exhaust bypass operation. The flow 
control element is positioned within the exhaust 
bypass conduit and the waste-gate actuator is respon- 
sive to bypass air pressure for actuating the control 
element. An air pressure control device has an intake 

port connected with the intake-air outlet and an outlet 
port providing bypass air pressure to the flow control 
actuator. When airflow paths between intake and 
outlet ports provides for continuous flow at least one 
other inhibits through airflow. 

Provision of a sprung element in the connect- 
ing rod of a piston engine is achieved by a series of 
belleville washers in the design of Fig 11. They 
constrain a piston within a piston, the inner one of 
which moves on a steel liner in the aluminium body 

Fig 10: Turbocharger waste- gate control by 
Cummins Engine (GB231 1556A) 



of the outer piston. A retainer ring at the base of the 
liner restricts the movement of the inner piston. 
Because, after ignition, the first element of the ‘stroke’ 
is compression of the springs and thus an increase in 
clearance volume, there is a drop in cylinder pres- 
sure, and temperature, in the combustion chamber. 
As the energy stored in the springs is released after 
TDC, although ignition occurs before TDC, engine 
efficiency is increased because there is no longer any 
negative torque occurring before TDC, changing the 
theoretical air cycle of the engine to a constant- 
pressure one. 

Engine induction systems 
Low-cost supercharging and 
electromagnetic valve actuation 

An inexpensive vehicle engine supercharger, having 
an integral by-pass mechanism, is proposed in the 
design of Fig 12. Of the rotary, positive-displacement 
type, it incorporates a duct within its casing to 
communicate with the suction and discharge ports at 
each end, for supplying naturally aspirated air to the 
engine. Rotors also compress the air entering from 
the inlet port on the right to supply the discharge port, 
to the engine, on the left. The by-pass passage is 
controlled by a valve and actuator on the extreme 
right of the mechanism. On the extreme left is the 
electromagnetic clutch which drives the supercharger 
but drive is broken when the by-pass valve senses a 
predetermined maximum boost pressure and opens 
the bypass port. 

Long valve opening time, and so adequate 
cylinder filling at high engine speeds, is among the 
advantages claimed for this electromagnetically ac- 
tuated poppet valve, Fig 13. The armature of the 
‘solenoid’ has switching magnets which hold the 
valve in open and closed positions. Compression 
springs are positioned between magnets and valve to 
avoid free-play in both the opening and closing 

Better fuel-spray atomization, 
integrated air/fuel induction system, 
fuel injection rethought and storing 
swirl energy 

The design in Fig 14 exploits the advantage that 

Fig 11: Constant-pressure cycle piston 
modification by GF Galvin (GB23181 51 A) 

Fig 12: Low-cost supercharging by Tochigi Fugi 
Sangyo (GB2301149A) 

Fig 13: Electromagnetic val\ 
actuation by Daimler-Benz 

Fig 14: Better fuel spray atomisation by Ford 



suddenly vaporized, or flashed, fuel effectively breaks 
up the spray from a fuel injection nozzle, when the 




71 ^ 7 

l -+- tr—- 1 / 









Fig 16: Fuel injection rethought by Robert Bosch 
( GB2295204A ) 

fuel is subjected to a rapid pressure drop as it dis- 
charges into the inlet port or combustion chamber. 
The proposed engine incorporates a heat-exchanger 
for pre-heating the fuel, upstream of the injectors. 
Temperature is maintained at a constant level just 
below the fuel's boiling point and the heating fluid 
may be coolant water or exhaust gas. Heating can be 
assisted by an electrical coil and insulated reservoir. 

Efficient engine air-intake silencing with mini- 
mum space usage around the engine is achieved in the 
design of Fig 15. The intake air-box and manifold 
ducts are integrated in one unit, the air-box volume 
surrounding the intake ducts. An air-plenum and 
throttle valve is also integrated into the assembly 
together with the fuel rail delivering to the injectors. 

A fundamental problem has existed with direct 
fuel injection into engine combustion chambers. This 
is in the conflict between providing a fuel-rich mix- 
ture for ignition at the sparking plug yet providing a 
narrow enough spray cone to avoid wetting the 
cylinder walls with fuel which would fail to vaporize 
and result in unwanted hydrocarbon emissions in the 
exhaust. In the design of Fig 16, a poppet valve type 
injector provides a narrow conical spray but has a 
subsidiary conical arc sprayed towards the sparking 
plug. This is achieved by a recess in the annular wall 
of the injector downstream of the valve seat. 

In an engine designed so that two intake port 
branches supply each inlet valve, the first branch 
being connected to a common plenum of an air 
supply manifold while the second is connected to a 
common plenum of a second manifold, this second 
manifold acts as a stratified charge storage arrange- 
ment as covered by an earlier patent by Ford. Now the 
company have extended this concept to lean bum 
engines so as to improve charge preparation and 
combustion quality. Fig 17. Air drawn from the first 
branch into the second swirls while in storage and 
retains kinetic energy for the next intake stroke. 
When the intake valve of the left hand cylinder is 
closed air is drawn through the lower throttle butter- 
fly into the dotted line drawn duct under the action of 
the intake depression caused by the adjacent right- 
hand cylinder as its valve opens, there being a tangen- 
tial swirl connection into the second plenum, down- 
stream of which fuel is injected and upstream of 
which the plenum duct is shaped to retain swirl 
energy. The upper of the two butterfly valves may be 
opened at high engine loads. 



Refinement and reduced 

More on the Sarich 2-stroke and a 
Rover K-series development 

The promise of greater fuel economy using two rather 
than four stroke engines has been given a new interest 
by the Australian Sarich engine which has caught the 
attention of the volume builders. Sarich has done 
much to advance the cause of two-strokes by allaying 
fears of poor emissions performance by devising the 
acclaimed system of injecting fuel-air mixture into 
the cylinder. The metering and injection system is 
shown in Fig 18. Fuel is metered through chamber (a) 
via port (b) and compressed gas is introduced via an 
air-port (c) to displace the fuel and deliver it to the 
engine. Crucial phasing of the valves is controlled to 
ensure fuel inlet and outlet (d) valves are closed 
before the air one opens and that any residual gas is 
expelled before the next charge. Gas inlet to the 
chamber is controlled by valve (e) opening at a higher 
pressure than it requires to close the outlets. 

A stretched version of the Rover K-series en- 
gine features variable valve timing in Fig 19. So- 
called ‘damp-liner’ technology has been used involv- 
ing bore walls just 3 mm thick and a water cooling 
jacket of just 0.65 mm. An extra long crank throw 
then brings swept volume to 1.8 litres. The VVT 
mechanism continually varies the inlet cam period to 

optimize power. An eccentric disc controlled by the 
EMS alters the phasing of camshaft to crankshaft so 
as to lengthen the inlet valve opening period above 
400 rpm. A maximum power of 143 bhp at 7000 rpm 
propels the 2359 lb vehicle (MG/ 7 ) at its maximum 
speed of 130 mph. 

Fig 18: Improved Sarich 2-stroke, in Patent 

Fig 19: WT for K-series 



Variable valve timing made easy, 
enhanced fatigue life for valve gear 
and variable geometry engine intake 

By using a lateral displacement device for the chain 
drive to an engine valve operating mechanism, a 
phase change can be obtained between the driver and 
driven sprockets in a system announced by Sachs 
Industries, Fig 20. Engine oil pressure is used to 
effect the displacement via a chain tensioner device. 
A check valve, shown in the sectional view of the 
tensioner, controls the operating pressure and a sole- 
noid valve connected to the pressure cylinder causes 
the device to change from chain tensioning to dis- 
placement mode. The solenoid is switched by the 
engine management system and in the arrangement 

Fig 21: Enhanced fatigue life 
for valve gear by Mercedes- 
Benz (GB2297805A) 

shown at A around 30 degrees of crankshaft angle 
displacement is involved in the phase-change. In the 
3 or 4 position arrangement at B, both camshafts can 
be controlled and, by using a proportional solenoid 
valve, a possible future application may be continu- 
ously variable cam-phasing. 

A valve drive system has been suggested which 
combines long fatigue life with maximum stiffness of 
all its constituent elements. In Fig 21 an IC engine 
camshaft is supported by a housing comprising pil- 
lars and beams which do not require complex ma- 
chining of the cylinder head. A short bearing interval 
ensures avoidance of camshaft bending vibration and 
the beams joining the bearing support pillars are only 
interrupted by cut-outs for the extra short rockers and 
thus form an unusually secure mounting for the 
rocker shafts. 

Obtaining induction ram effect at more than 
one ‘resonant’ speed is the object of the design in Fig 
22. The carburettor or fuel-injection choke tube is 
connected to a plenum chamber in a barrel rotatable 
by a motor geared to engine speed. The effective 
lengths of the intake and outlets vary with barrel 
position by their formation, in part, by grooves in the 
barrel wall. 











>j FQ^t on jY ^TffTl g 


v — jV 









y back .'mamo-:* 


[XHAUSI camvum 

A g . , .r 

NlAKl UnSHAfl 



H) deyrm 

Fig 20: Sachs variable valve timing 

Fig 22: Variable geometry intake by Tickford, in 
Patent GB2205609A 



Economical flywheel to crankshaft 
connection and secondary engine 
force balancer for motorcycle 

A more cost-effective connection between engine 
flywheel and crankshaft is possible, without loss in 
reliability, with the design in Fig 23. It uses a collar, 
held by an expansion screw, in a concentric opening 
in the crankshaft. Dog teeth on shaft and flywheel 
form the connection, the joint being made firm by 
tensioning of an axial spring acting on an expansion 
screw where the two parts mate. The flywheel has a 
collar which extends into the crankshaft opening, the 
centre of the collar having a hole through which the 
expansion screw operates. 

Normally there is difficulty incorporating a full 
Lanchester-type harmonic balancer into a motor 

involved. In the design of Fig 24 , a compromise is 
reached in which a large, crankshaft-mounted, gear- 
wheel drives two smaller co-planar gears in opposite 
directions, one by a further countershaft gear. 

Low emission CVs 

Initiatives taken by MAN for diesel engine pollution 
reduction include the air-injection system shown in 
Fig 25 for dealing with visible smoke on initial take- 
off and subsequent acceleration before the turbo- 
charger achieves full compression. Normally fuel is 
not completely burnt under these conditions as the 
engine is not supplied with enough air — so an air 
reservoir is introduced which is fed from the com- 
pressor. The injection system comprises a pressure 
pipe with solenoid valve and non-return valve in the 

cycle engine because of the high rotational speeds charge-air line at the inlet to the air distribution pipe. 

i By means of an electronic controller, throttle pedal 

depression opens the solenoid valve to inject air for 

R^cvt:l' r *n htrrn*v 

3o***o«c» AX 
valv* M 

V C<nno«^5S<*<l a*r r 

Fig 23: Economical Fig 24: Balancer 

flywheel to crankshaft system by Triumph 

connection by Fichtel & Motorcycles in Patent 

Sachs (GB2301166A) GB2186914A 

Fual injection quantity j 
• (Control rod travel) I r _ 

Accelerator pedal 

Air Infection 

Control •laeironiea j. . . . / T 

I 1 I : 

Particulate filter 

Electric ’ff 
motor : Filter 

Fig 25: MAN diesel pollution control 



a 0.7 second period. Closure of the non-return valve 
protects flow from the turbocharger while the ECU 
releases a matching quantity of fuel for injection. The 
system is used on the company’s NL202 and NG272 
low-floor buses. The company has also developed its 
particulate filter technology in which both duplex 
and full-flow arrangements are used to overcome the 
problems of filter clogging with deposits which can- 

Monitor lamp 

Fig 26: Filtering diesel particulates by Mercedes- 

Fig 27: Speeding engine warm-up, by Ford, 
covered in Patent GB2199368A 

Fig 28: Quicker catalyst light-up time by Rover 
Group (GB2316338A) 

not be burnt off in conventional layouts because 
exhaust gas temperature from low-consumption en- 
gines is too low. In addition, oxidiszing catalytic 
converters are being developed for exhaust gas post- 

An exhaust filtering system for diesel 
particulates is now offered on Mercedes bus chassis. 
The so-called ceramic candle type filter features 
catalytic regeneration. The system, Fig 26, complies 
with emission limits for urban buses which came into 
force in the US during 1991 . The company is reported 
to have rejected the philosophy of developing cata- 
lytic control systems originally engineered for spark- 
ignition engines, on the grounds that diesel emission 
control needs solutions with less flow resistance, 
proneness to blockage and better thermal and me- 
chanical resistance in the harsher CV environment. 

Speeding engine warm-up, quicker 
light-up time for denox converter 
and use of scotch-yoke crank 

Exhaust ducting cast in the cylinder head is used in 
this design proposal. Fig 27. The ducts allow alter- 
nate paths for the engine exhaust according to posi- 
tion of two-way valves. As well as warming up the 
inlet manifold during initial running of the engine, 
exhaust gases also warm up the engine coolant and 
lubrication systems. 

Faster catalyst warm-up time from engine start- 
up is the objective of the design of Fig 28 which 
involves providing excess fuel and air to enter the 
converter for part of the combustion cycle. A fuelling 
device controls the exhaust gas entering the catalytic 
converter such that a combustible mixture is intro- 
duced to enable the converter to more quickly reach 
light-off temperature or regain it if very cold operat- 
ing conditions cause the catalyst to ‘go out’. 

The Collins CMC scotch-yoke engine, Fig 29, 
has attracted considerable interest. Exhibited re- 
cently by consultants Emdair who are R&D contrac- 
tors to the Australian builders, the engine has in- 
verted-tee cylinder formation in the 2-litre two- 
stroke version shown here. Compact size is thanks 
mainly to the scotch yoke mechanism joining the 
pistons to the crankshaft. It has been calculated that 
a vehicle designed to exploit the lower package size 
of the engine would achieve a 10% fuel saving over 
a conventional design. 



Drive and steer systems 
Transport of the future — guided 
lanes or electric/flywheel hybrid car 

Tribus, Fig 30, has been described as being to the 
trolleybus what light-rapid-transit is to trams. But it 
also applies to both goods and passenger vehicles — 
to urban and inter-urban traffic. A segmented bar 
pantograph collector overcomes the traditional limi- 
tations of trolley booms. This results in a multi-lane 
high speed equivalent to a trolleybus which could 
have attractions both for commuting and dense traf- 
fic motorway driving. The overhead wiring structure 
is vastly simplified from traditional trolleybus lay- 
outs and the ability to ‘jump’ from lane to lane, and 
through junctions, is made possible by reserve energy 
provided by on-vehicle flywheels and/or hybrid elec- 
tric/petrol drive systems. On the three lane motorway 
installation shown, only the two inside lanes would 
typically be electrified. Inset is shown how two or 
more elements of the segmented pantograph bar are 
always in contact with overhead feed cables. The 
originators are Lawrence Designs, 40 Lower Broad 
Street, Ludlow, SY8 1PH. 

The US company American Flywheel Systems 
is claiming ranges for electric cars of 300-600 miles 
against the 50-100 miles which is representative of 
existing battery-electric technology. The AFS ‘me- 
chanical battery’, Fig 31, stores kinetic energy in 
vacuum-enclosed rotors running at high speeds on 
magnetic bearings. The charge current causes the 
rotors to accelerate up to speed and energy is drawn 
from the spinning windings by the unit acting as a 
generator. The technology is made possible by new 
materials for the fibre windings and the magnetic 
bearings. The vacuum housing is shown at (a), fibre 
windings at (b) and stationary shaft at (c); counter- 

A J 

Fig 31: Flywheel battery 

rotating rim wheels are at (d), spoke tubes containing 
the magnets at (e) and magnetic bearings at (f) — (g) 
and (h) being the pick-up coils and end-bell respec- 

Fig 29: Collins scotch-yoke engine 

Fig 30: Guided lane system 



Fig 32: Sure-fire gear engagement by Mercedes- 
Benz (GB2308165A) 

JCB Transmissions 

Ensuring sure-fire gear engagement 
and muiti input/output gearbox 

Inertial effects which might impair proper gear- 
selection in a change-speed linkage, between shift- 
lever and selector forks, are negated in the design of 
Fig 32. An extrusion lever with an inertial mass at its 
end is linked by a ball-jointed pull-rod to the gear- 
change lever on the selector shaft. The geometry of 
the linkage is such that pivot arm of the ball-joint, on 
the extension-arm, to the selector shaft is larger than 
that between the ball-joint and the arm’s own pivot- 
point. Also the pivot arm between the ball-joint and 
selector shaft axis on the change-gear lever is larger 
than that between the hinge-point which connects to 
the shift-lever, and that lever’s pivot-point, on the 
selector shaft axis. 

An additional transmission path, providing 
further gear ratio, is the objective of the design of Fig 
33. The first input A has the conventional output 
while a second input B has a mechanism C associated 
with it, and a common gearwheel D, which allows 
incorporation of part of the change-speed train of the 
main gearbox to be used with the secondary transmis- 
sion path which has output to a transfer drive at the 
base of the assembly. 

Rapid actuation diff-lock, improved 
racing clutch and automatic drive 

The differential housing in Fig 34 is turned by the 
final drive gear and contains a pair of auxiliary drive 
gears, engaging their respective pinions. These rotate 
in bearing apertures within the housing. A lockable 
part A is coupled to a differential part, tumable in one 
direction with respect to the housing. A locking 
element B also turns with the housing and a coupling 
mechanism C is located between locked and free 
elements to block rotation of the differential, oper- 
ated by the adjustable member D. A cam mechanism 
E increases the coupling force. 

Avoiding undue wear in the slipping of high 
performance car clutches is the objective of the 
design in Fig 35. The main, driver-disengageable, 
clutch is supplemented by a smaller, belleville-spring- 
loaded, clutch between the drive plate and the splined 
gearbox input shaft. The second clutch is arranged to 
slip at a higher torque than that of the main clutch, 



which is of a multi-plate carbon-carbon type. Be- 
cause the high torques which cause slipping of the 
secondary clutch are only applied for brief periods, 
little heat is generated. 

Power transmission to the front or rear axle of 
a vehicle can be made a function of the speed 
difference between the front and rear wheels but, thus 
far, the necessary mechanisms have been unwieldy. 
In the design of Fig 36, a compact drive splitting 
mechanism is devised offering considerable freedom 
of location on the underside of the vehicle. A trochoidal 
gear pump is mounted between a shaft and co-axially 
rotating cylinder, each connected respectively to 
front and rear axle prop-shafts. A piston slides in the 
cylinder in response to pressure generated by the 
pump to actuate a clutch which locks front and rear 
drive together. 

Easy-change CV-UJ assembly, 

Better performing plunge joint and 
better CV-joint packaging 

Ease of replacement of constant-velocity universal 
joint assemblies is considerably improved by making 
the hub and joint housing members as separate units. 
Fig 37. The hub comprises a flange portion, to which 
the road-wheel is bolted, and an integral body sup- 
ported by hub-bearings for which the body forms the 
outer race. Hub and joint housing each have interfac- 
ing conical surfaces as well as an interconnecting 
splined joint. A spigot on the joint housing has a 
projection on it which is deformed on assembly so as 
to lock on to the conical surface of the hub. 

Attempts have been made to reduce the plunge 
resistance of tripode universal joints under high 
installed angles by the use of both inner and outer 
rollers mounted to the tripode trunnions, each having 

Fig 38: Better performing plunge joint by NTN 
Corporation (GB2299393A) 



spherical outer surfaces and cylindrical inner ones. 
Although induced thrust is lower than in earlier types 
further reduction is limited by the tendency of the 
outer rollers to follow the inner rollers because of 
friction between them with increase in both contact 
stresses and rolling resistance. In the arrangement of 

Fig 39: Better CV -joint packaging by Lohr and 
Bromkamp (GB2295365) 

Fig 40: Balanced torque triaxle drive 

Fig 41: Sophisticated auto-steer for trailer by GW 
Mitchell (GB2315471A) 

Fig 38, the inner surfaces of the outer rollers are such 
as to direct a load component parallel to the axis of 
each trunnion. 

By offsetting the centre of the CV joint out- 
wards of the centre plane of the front wheel bearing, 
several advantages have been achieved in this inven- 
tion, Fig 39. A longer drive shaft, resulting, means 
less angularity on bump and rebound. Since the inner 
tracks of the wheel bearing lie axially adjacent to the 
outer running grooves in the outer member of the CV 
joint, bearing performance is unaffected by torque 
transmitted and the overall bearing diameter can be 
reduced. Finally the CV joint can be assembled from 
the outside into the hub member, which doubles as 
outer race of the CV joint and inner race of the wheel 

Balanced torque triaxle drive, 
sophisticated auto-steer for trailer 

In order to overcome the problems of uneven torque 
splitting between the three axles of a triaxle drive, the 
design in Fig 40 by GKN Axles proposes the use of 
a spur gear epicyclic differential unit at the point of 
torque input to the bogie. This is designed to split 
torque unevenly, passing some 67% of the input 
torque down a through shaft inside another tubular 
shaft which conveys the remaining 33% to the first 
axle inter-wheel differential unit. The initial 67% can 
then be split evenly between second and third axles 
in the normal manner. 

Automatic steering of the leading and trailing 
axles of a three axle trailer bogie is possible, in the 
design of Fig 41, in response to the relative move- 
ment of the main chassis with the bogie suspension 
sub-frame, a change in direction of the steering being 
effected with just one movement of the steering 
mounting plate. The movement of the steering mount- 
ing plate, relative to the main chassis between its 
outermost positions, is such that the line of action of 
the joining-ends of the primary link crosses the axis 
about which the main chassis rotates. The leading 
and trailing axles are pivotable with respect to the 
triaxle subframe, a linkage mechanism coupling each 
of the steerable wheels to the steering mounting plate 
on the main chassis. The mounting plate rotates 
between two positions which govern the direction of 



Wheeled tanks challenge tracklayers 
and more efficient engine-coolant 

A challenge to tracklaying armoured fighting vehi- 
cles was made by Timoney Technology of Ireland in 
a presentation to Autotech, Fig 42. Arguing against 
the assertion that wheeled vehicles are more vulner- 
able than tracked ones in a combat situation, the 
authors argued that an 8 X 8 off-roader could often 
keep going with one wheel station out of action. A 
notable feature is overall height of just 1 .3 metres to 
the turret ring, some 500 mm lower than other 
vehicles of this category. Axles are fitted with in- 
board hydraulic disc brakes and have differential 
lock facility. Patented wheel hubs have a one-piece 
stub-axle casting with lugs to carry the wishbone 
ball- joint bearings. The planetary gear system is also 

inboard of the stub axle and the ring gear is attached 
to and rotates with the hub housing — importantly 
permitting a small diameter rotary air seal for the 
centralized tyre inflation system. Suspension is of the 
double-arm type with two springs mounted on the 
lower wishbone, resulting in low overall profile and 
minimum hull intrusion. Suspension movement of 
400 mm is supplemented by high deflection of the 
large tyres. Maximum speed for the 24-31 tonne 
combat weight vehicle is over 100 km/h. 

Reduction in back-flow of air drawn through a 
vehicle engine coolant radiator, by the fan, is the 
objective of the design in Fig 43. The cowl which 
extends backwards from the radiator, to the fan, 
incorporates a C-shaped scoop facing the clearance 
around the fan. Back-flow is thus inhibited by turning 
it back on itself to create eddies within the fan-blade 
tip vortex region. 

Fig 42: Wheeled tanks challenge tracklayers 

Fig 43: More efficient engine-coolant airflow by 
Rover Group (GB2311562A) 



Suspension development 
Transverse spring composite 
suspension and optimizing 
unsprung mass with wheel travel 

A transverse single leaf spring made from polymer 
composites forms the lower control arms for the 
wheels and hubs in the suspension system design of 
Fig 44. Where the spring attaches to the central 
mounting it is in the form of an inverted dihedral and 
remains clamped as such even after the normal load 
of the vehicle has deflected the outer arms of the 
spring into an upwardly inclined attitude. 

In the suspension configuration for driven 
wheels of Fig 45, part of the mass of the transmission 
unit is supported as sprung mass by the body. At the 
same time, vertical movement of the transmission is 
allowed so that the wheels can also travel vertically 
over an increased distance before the universal joints 
reach their angular limits. The suspension arm for 
each driven wheel is connected to both sprung and 
unsprung parts of the vehicle while drive to the 
wheels is through articulating shafts. During motion 
of the suspension, vertical motion of either of the 
wheels will be transmitted through the respective 
arm to produce a smaller vertical movement of the 
transmission unit. 

Advanced suspension linkage 

Active rear axle kinematics is the name given by 
BMW for its rear suspension design on the 850i 
coupe, Fig 46. As well as providing anti-squat and 
anti-dive, wheel-camber is adapted to reduce tyre 
strain at high speed and roll-steer is built-in which 
achieves load reversal reactions. As well as four 
wheel control links, a fifth joins two of the others to 
modify their relative motions. Two lower and one 
upper transverse link form the ‘wishbones’ while a 
trailing link provides fore-aft location. This link has 
a secondary connection via an intermediate link to 
the upper transverse link. Coil-spring and damper are 

covered in Patent GB2202498A 

Fig 45 : 
unsprung mass 
with wheel 
travel by Rover 

(GB2 302066 A) 

mounted adjacently on the rear wishbone. The wheel- 
carrier pivot is above and ahead of the wheel-axis, to 
achieve anti-dive/squat. The intermediate (fifth) link 
is a strut which transfers braking torque on the wheel 
carrier, through the upper ‘wishbone’ to the trailing 
link. Because the strut lower pivot has less velocity 
than the wheel-carrier pivot on the trailing arm, 
during vertical travel of the wheel, this velocity is 
also seen at the upper wishbone. This causes the 
wheel carrier to rotate more slowly than the trailing 
link about the wheel-axis, and twisting of the wheel 
carrier on braking is suppressed without losing the 
horizontal compliance in the anchorage bushes of the 
diff-carrier at the root of the transverse links. The link 
also compensates for horizontal forces on the wheel 
either side of the vehicle so no swivelling of the 
whole suspension, with the diff-carrier, takes place. 
The elasto-kinetic function arises from the interac- 
tion of the anchorage bushes with the linkage geom- 
etry. In braking, for example, the two links forming 
the lower wishbone are subject to differential tension 
and compression. At the same time the anchorage 
bush of the trailing arm yields. Transverse link 
motions are such as to ensure that resulting motion of 
the wheel carrier is purely horizontal, without caus- 
ing steer effects as the wheels move vertically. With 
the short wheelbase of the coupe compared with the 
saloon, and the comparatively high driving torques, 
smooth behaviour under load reversal conditions has 
to be ensured. This is helped again by the design of 
the trailing arm anchorage. At high lateral 
accelerations, wishbone links are distinctly skew to 
the body of the tilting vehicle. With load reversal 
caused by sudden removal of driving torque, the diff- 
carrier swivels due to the resilience of its mounts. 
Simultaneously it puts the pivot point of the upper 
wishbone link within it, horizontally, and the front/ 
rear lower wishbone links are displaced by (fv) and 
(fh) respectively. Tilt of the wishbone also causes 
lateral displacements of the pivot points (rv) and (rh), 
causing understeer in both wheels. This effect is only 



Improved suspension geometry 
control, avoiding suspension 
compliance conflict/steer fight in 
braking/accelerating and smoother 
action for suspension strut 

The desire to change from rigid beam axle to inde- 
pendent suspension on vehicle categories such as 
4X4 sports/utility vehicles may be restricted by con- 
cerns for loss of accurate suspension geometry, par- 
ticularly when an off-centre differential is required to 
clear an engine sump and unequal length drive-shafts 
are involved. In Fig 47 a compact form of IFS is 
proposed. Tubular members which surround the drive- 
shafts are pivotally connected to the hub carrier units 
and chassis-mounted differential casing in a ‘wish- 
bone’ arrangement. The wishbone length is maxi- 
mized, by the position of the pivots, to give better 
geometric compromise. 

A torsion bar front suspension system for a 
vehicle which is simple in construction, and yet does 



Automotive ( GB2295988 ) 

not result in bending of the torsion bar when the 
vehicle is subjected to lateral loading, has been 
achieved in the design of Fig 48. By ensuring the 
torsion bar axis is at a small angle to the axis of the 
lower suspension arm fulcrum and laterally dis- 
placed from it, appropriate values of La, Lb, Lc, S 
and 0 can be chosen so that sufficient suspension 
compliance is available in the fulcrum bushes with- 
out any bending moment being applied to the torsion 
bar. The spring plate connecting the bar to the arm is 
slender enough to be flexible in the fore-aft direction. 

Reduction of the steer effect in braking is 
claimed for this suspension configuration. Fig 49, for 
the steered wheels of vehicles. Compliant bushes, B 
and B , normally permit the suspension arm to rotate 
about an axis A as the road wheel rises and falls but 


this can change to either A or A under appropriate 
loading. During braking or acceleration the angle 0 
and the compliance of the bushes allows the suspen- 
sion arm to mm in the horizontal plane and compen- 
sate for steering displacement of the road- wheel hub. 

Reduction in ride harshness is achieved in this 
independent wheel suspension linkage configura- 
tion, Fig 50. The centre portion of the spring-damper 
stmt is used as an anchorage for the anti-roll bar, a 
link from which is angled to the axis of the suspension 
stmt in such a way that transverse reaction loads from 
the ground-level cornering forces are balanced by 
loads induced in the anti-roll bar so that sideways 
deflection of the damper tube-and-rod is minimized. 


Fig 49: Avoiding steer fight in 
braking/accelerating by Rover 
Group ( GB2311499A ) 

Fig 50: Smoother action for 
suspension strut by Rover Group 
(GB230641 1 A) 



Fluid-actuated anti-roll control, non- 
stick front forks and freer moving 
bogie suspension 

Lower weight and greater compactness is said to arise 
from the use of this fluid-actuated anti-roll system, 
Fig 51, which claims to achieve a high degree of 
lateral stabilization. Hydraulic actuators act on the 
upper ends of telescopic suspension struts; the actua- 
tor pistons are hydraulically linked in a way that 
motion in one is mirrored in the other. Excessive load 
on one suspension unit is thus transferred to the 
opposite side unit varying according to vertical sus- 
pension travel. 

Overcoming the problem of high friction in seal 
systems used on the front forks of motor cycles is the 
subject of the design in Fig 52. The usual telescopic 
fork arrangement gives way to a twin-wishbone sus- 
pension arrangement. Universal joints at the outer 
ends of the wishbones allow steering motions as well 
as vertical suspension travel. A pivoting triangular 
framework is also used to transfer steering torque 
from the handlebars to the forks which support the 
front wheel. 

Less chance of bearing seizure at opposite ends 
of the walking beams of a tandem bogie suspension 
results from this design, Fig 53. Forked axle brackets 
connect to the beam ends via exposed bolts with 
bushes and convenient stroke mechanism — which 
also lowers the vehicle to the ground on electric jacks. 
Steering involves the use of a highly flexible steering 
wheel including a fork or ball used to give the user a 
stable support. A voice command system is used for 
a number of the secondary controls. 

Fig 51: Fluid-actuated anti-roll control by AMG 
Motorenbau & Entwicklung (G B2315716A) 

Fig 52: Nonstick front forks by NH Flos sack 
( GB2207645A ) 

Fig 53: Freer moving bogie suspension by the US 
Boler Company in Patent GB2192596A 



Fig 54: Easy prestressing for swivel pin bearings 
by Claas KGAA (GB2316050A) 

Fig 55b: Steering column immobiliser by 
Marshall Wolverhampton (GB2303108A) 

Fig 56: Tiller steering by Kazunaga Morimoto in 

Easy prestressing for swivel pin 
bearings, uncomplicated 
steering-castor adjustment and a 
steering column immobilizer 

An unusual arrangement of steering swivel pins on 
this steer-drive axle for industrial vehicles, Fig 54, 
allows prestressing of the taper roller bearings to be 
carried out without difficulty. Outwardly facing swivel 
pins are used in a design which is said to be less 
expensive than normal arrangements of comparable 
performance. Forked ends of the axle body, contain- 
ing the pins, engage with forked ends of the hub 
carrier. Each swivel is in the form of a circular pin 
provided with central collar and constrained at one 
end in a bore of the axle body fork. One of the 
corresponding bores in the hub carrier fork has an 
internal thread into which a cover is screwed with an 
inner collar bearing against the outer race of the taper 
roller bearing. 

The problem of providing degrees of castor 
adjustment without weakening the integrity of wheel 
location, or involving heavy expense, is alleviated in 
the design of Fig 55. A steering ball joint incorporates 
a wide flange which overlaps the transverse link to 
which it is connected and there is a connecting dowel 
which pivots one part with the other. The parts are 
clamped together through cleverly profiled holes for 
the securing bolts. These allow three different angu- 
lar positions between the joint and the link so as to 
provide stepped adjustment of castor with secure fit 
when the clamping bolts are tightened. A compara- 
tively modest torque can be used to lock a steering 
wheel/column against rotation by using a high ratio 
gear pair between the column and the clutched lock- 
ing device. Fig 55b. A member which has coded input 
response to a key or other immobiliszer controls the 
action of the clutch, which is based on teeth cut into 
the shaft of the pinion of the gear pair. 

Back to tiller steering and an anti- 
dive motor cycle suspension? 

As the concern for driver safety increases, the very 
concept of a circular steering wheel attached to a 
rigid column will be questioned. This is the view of 
an invention. Fig 56, which replaces the wheel with 
a laterally moving handle shaped to optimize protec- 
tion of the thorax region, on impact. The handle is 
connected to a rack-and-pinion by deformable arms. 



and an hydraulic valve, which transfer the side-to- 
side steering motion into swivelling of the road 

Two fore-and-aft suspension arms form a four- 
bar linkage which gives anti-dive suspension geom- 
etry to the front wheel of a motor cycle when braking. 
In the arrangement of Fig 57, geometry is such as to 
allow positioning of a steering arm which articulates 
without steer fight on vertical travel of the suspen- 

Control strategy for 4-wheel steer, 
self-steering in forward and reverse 
and a necked hub member 

Lotus have released more information about the 
active rear steer system they have under development 
— which they say is an extension to their active 
suspension work using closed loop control strategies. 
This allows steering to respond to both driver inputs 
and external disturbances. Vehicles thus equipped 
would therefore respond to crosswinds or split brak- 
ing in a controlled manner. There is also less sensitiv- 

To overcome the limitations of existing self- 
steering axles, this design by an Italian inventor, Fig 
59, comprises a tubular member which slides over the 
axle while road wheels pivot directly on the frame of 
the vehicle. Two tie-rods positioned on either side of 
the axle have their ends linked eccentrically to the 
wheel hubs. Thus, while reversing or moving for- 
ward, actuators are caused to restore the tie rods and 
therefore the wheels to their neutral position. 

Improved heat insulation for the tyre and greater 
space for the brake actuator cam arise in this hub 
design. Fig 60. The tapered hub section and necking 
between the bearings produce the desired configura- 



E-Booster Modified Basic Unit LSC 80 (10") 
Master Cylinder 22.2 C V/CV 
Function I A utomatic Braking 

Function II Automatic Booster Operation 

Input Rod Force 
Input Rod Speed 
Kneepoint 80 bar 

Response Time 200 ms 

Release Time 180 ms 

Function III Variable Ratio - Pedal operated 
I nput Rod Force 1 400 N 

Input Rod Speed 0 - 100 mm/s 

Kneepoint 1 25 bar 

Response Time 1 70 ms 

Release Time 1 80 ms 

Fig 62: Initial booster + proportional valve 



Fig 63: Variable ratio characteristic (full-line 
base ratio; dashed line medium ratio; chain line 
high ratio) 

Braking systems 

Electronic brake actuation system 

Lucas Varity have determined that the typical driver 
is usually inhibited against applying optimum pres- 
sure to effect an emergency stop. The company are 
therefore advocating the use of an electronic actua- 
tion system (EAS) based on the electronic booster 
seen in Fig 61 that has been replacing the fast- 
vacuum booster, fitted to many cars, and which was 
limited to providing assistance only at a fixed ratio 
and often reacted too slowly in emergency. 

A proportional control valve has been devel- 
oped such that the output stroke is proportional to the 
input control current signalled through the brake 
pedal. A control loop is partially closed via an ECU 
which measures output/input force ratio and com- 
pares it with a vehicle-specific algorithm. An initial 
specification was as shown in the table of Fig 62 and 
involved electronic booster control superimposed on 
a mechanical booster. Performance curves for a 
series of possible ratios are in Fig 63. 

A parallel development of the long-stroke (LS) 
booster meant that the benefits of electronic control 
could be applied to heavier cars and replace tandem 
booster arrangements which tend to be restricted to 
either high-performance cars or heavier van deriva- 
tives. The company is concentrating on the develop- 
ment of the LSC 1 1 5 T tandem power unit, Fig 64, 
which has an integrated electronically proportioning 
solenoid control valve. Shown enlarged in Fig 65, 
this valve is seen to be an electrical analogue of the 
proven servo control valve, but with input/output 
ratios compared electronically rather than by rubber 
reaction disk. 

The company believe the full potential of such 
systems can only be realized when legal insistence on 
mechanical back-up is dropped and 100% reliability 
of the electronics can be guaranteed. Money saved 
could be used to incorporate other electronic fea- 
tures, Fig 66, such as ABS and engine management. 
The ECU would be able to compare actual vehicle 
deceleration with the theoretical value based on a 
given booster output force, and thus forewarn of 
brake-fade. The ECU could also relate engine torque 
to achieved acceleration for assessing loading on the 
vehicle so that the system is informed of prevailing 
road gradient. In integrating ABS, the booster would 
provide the energy source and the combined system 



ECUs could help to create the next stage of automatic equate. For stage three the problem of varying brake 

braking, in controlled traffic conditions, involving pressure requirement with degree of vehicle payload 

intelligent cruise control. Functions such as traction is tackled. Here, not only is brake booster pressure 

control, hill-hold and variable pressure boost for metered but also the pressure level boosted propor- 

ABS could be incorporated. tionally to the pedal force (measured by a sensor on 

The first stage of the project will be seen on the piston rod) applied, so that constant deceleration 

some Mercedes cars early in 1996 as 'brake assist- is obtained with a constant pedal force irrespective of 

ant’. For this application the vacuum chamber of a 
conventional vacuum booster is equipped with the 
Mercedes position sensor. This tells the ECU that the 
booster piston has travelled a given distance in less 
than a pre-defined time and switches on the solenoid 
valve. Atmospheric air then enters the booster work- 
ing chamber to amplify driver effort with maximum 
available servo power. Without this device full 
decelerative advantage with ABS is lost. In an active 
stability control feature, the ECU helps to control 
side-slipping out of a comer by selectively applying 
braking force to one of the outside wheels, generating 
braking force from the booster and modulating it by 
the ABS. In stage two, hill-hold will be added such 
that brake pressure will automatically be applied 
when a vehicle stops on a hill and automatically 

released when it pulls away. This will involve the \ E° u f° r Sol, ‘ nold 

proportional solenoid valve and the booster generat- 3 Functional switch 

ing the necessary braking pressure by metering the air Fig 64: LSC 115 T with integral solenoid valve 

intake into its working chamber. Such metering is 
also essential for automatic braking and adaptive 
cruise control, ACC. Current cruise control, it is 
argued, can suffer a lack of sufficient engine decel- 
eration on steep gradients such that the vehicle gains 
speed; EAS can supply precisely metered braking 
pressure to prevent this. Beyond this, EAS can gen- 
erate pressure for ACC so that a vehicle is kept at a 
safe distance from the one ahead, without having to 
rely on engine-braking which might prove inad- 

E toe Ironic Actuation Syatam 

| EAS 3 | 

Booster with proportional Sohnokt Volvo 
and Input Output Signal* 

Energy Supply lor Automatic Brake 
Application and variable Sootier Ratio 
_ functions 

9 Active Dover Support 
9 Traction Control 
I 9 Car Dynamic Control 

Hill Hold 



TC - Brake 

TC - Brake 

TC - Engine 





TC- Engine 

HW Hold 

Fig 66: Features of EAS 3 with (right) interaction of ABS, EAS and EMS 



Fig 67: Brake apportioning for solo or coupled 
mode by Daimler-Benz from Patent GB2146720A 

Fig 68: Electromagnetic 
braking by Siemens 

Fig 72: Brake booster f GB2305478A ) 
by Delphi France 
( GB2312718A ) 

Fig 71: Failsafe brake actuator by Mercedes-Benz 

Brake apportioning for solo or 
coupled mode, electromagnetic 
braking, auto-adjusting electro- 
brake and fail-safe electric brake 

A car braking master-cylinder is ‘tandem-like’ in 
configuration and the pressures in front/rear wheel 
brake chambers apportioned according to vehicle 
load. This effect is altered according to whether or 
not a trailer is coupled, in the design of Fig 67, by 
means of a 3/2 way solenoid valve. 

A means of applying braking selectively to 
individual wheels is provided in the design of Fig 68. 
The brake actuator is rigidly mounted on the disc- 
brake calliper and uses electromagnetic operation. 
The actuator comprises an electric motor, the arma- 
ture spindle of which is threaded so as to produce 
translatory motion/force against the pad pack-plate 
for rotary motion of the armature. 

An electrical brake actuation system is covered 
in the design of Fig 69 which has the advantage that 
braking fluid does not have to be relied upon to 
balance out brake-pad wear. The electromagnetic 
actuator comprises a spindle axially driven by an 
electric motor via a threaded roller drive with the 
axial force boosted by a lever mechanism in transfer- 
ring it to a pressure plate acting on a brake piston. A 
mechanical correcting device is incorporated for 
pad-wear balance. 

An indication of the applied braking force is 
supplied to the driver in this electric, or electro- 
hydraulic, brake actuator which is backed-up by 
hydraulic brake application upon detection of fail- 
ure. A sensor behind the brake pedal detects any 
electrical faults and triggers the adjacent hydraulic 
system into actuation, while providing a reaction 
force at the pedal. The hydraulic unit has a through 
bore, to the master cylinder, containing two pistons 
which can slide relative to one another, Fig 70. 

Only a relatively small movement of the brake 
pedal is necessary in the design of Fig 7 1 , in the event 
of hydraulic pressure failure, before mechanical cou- 
pling of piston rod and piston takes place within the 
master cylinder. Mechanical coupling is achieved 
between two co-axially sliding members when the 
length of the piston stroke exceeds a threshold value. 
A locking catch, at A, is spring-loaded in the open 
position normally but is closed by the action of a cam 
when the stroke exceeds the threshold. 



Easy-assemble brake booster, 
compact integration of clutch servo 
actuator parts and set-speed down- 
hill braking 

In the Fig 72 design a booster housing having front 
and rear walls is held together by through tie-bolts 
which also form the mountings to a vehicle bulkhead 
and a means of securing the brake master cylinder to 
the housing. At the master-cylinder side of the booster, 
the outer wall of the housing is assembled against a 
splined collar which also gives the booster a measure 
of controlled collapse on impact. 

A compact clutch actuator, which can be easily 
installed, is the essence of the design of Fig 73. 
Pressurized fluid is used to actuate the clutch and the 
device incorporates a pump and drive, to the left of 
the diagram, feeding a pressure reservoir within the 
fluid container, to the right of the diagram. There is 
a valve, shown centre, between it and a servo cylin- 
der. Between reservoir and container there is a preset 
passage-constriction formed, by their integration, to 
restrict and control fluid flow. 

Problems of automatic down-hill ‘active’ brak- 
ing have been identified as due to uneven axle 
loadings due to the shift of vehicle centre-of-gravity 
on the incline. This causes wheel-locking in the 
lightly laden axle and consequent undue use of the 
ABS. In the design of Fig 74 the braking effort 
distribution is such that only the highly laden axle is 
braked in normal cases. Only when deviation from 
the set speed or slip at the braked wheels exceeds a 
certain value will the brakes be applied on the lightly 
laden axle. 

Easy height-adjust for trailer 
coupling and auto-adjusting towbar 

Existing adjustable-height trailer couplings are un- 
wieldy to adjust because of the heavy knuckle forgings, 
incorporating the hirth gears, being awkward to re- 
move and replace for resetting the engagement of the 
gears. In Fig 75 a parallelogram linkage is formed out 
of the front and rear supports of the coupling. The 
hirth gears are at the lower end of the front support and 
a gas spring damps the movement of the mechanism 
during adjustment. The geometry is such that up to 
half of the total towing force is taken up by the rear 

Loss of payload carrying area on a drawbar 
trailer truck combination arises from the need to have 

an inordinately long drawbar so that prime-mover 
and trailer do not foul one another on sharp cornering 
manoeuvres at low speed. In the design of Fig 76, the 
drawbar is made to increase in length on acute angled 
comers and in normal driving can be set much closer 
than is usually the case, to the advantage of available 
load space within a maximum overall length of 
combination. A tractor to trailer angular measuring 
device on the tow bar has an arm that engages a stop 
displaced from the axis of the towing hook. This 
signals a control unit which extends a hydraulic 
cylinder in the telescopic drawbar. 

Fig 74: Set-speed braking by Wabco GmbH 

Fig 75: Height-adjust coupling by Bradley 
Doubledock (GB2318107A) 

* ■ \ 

Fig 76: Auto-adjust towbar by Fruehauf Corporation 

Chapter 7: 

System developments: body structure/systems 

A decade of patents specifications and company releases 
provides an insight into automotive product development up 
to the period of new designs covered in the first five chapters. 

In the area of body structures and systems, the trends to 
reducing vehicle weight and increasing occupant safety are 
clear in the examples which follow. 




Controlled collapse 

Safety and weight reduction continue to be the main 
drivers of structural developments. Better control of 
frontal collapse is receiving particular attention. 
Avoidance of bending collapse of the front longitu- 
dinal members of a vehicle body in impact is the 
objective of the design in Fig 1. Because such mem- 
bers usually involve a curved section to provide 
clearance with mechanical systems it is difficult to 
prevent the onset of bending collapse, under end 
load, prior to the desired controlled longitudinal 
collapse of the box sections. While vertical beads are 
formed into the vertical walls of the box-members to 
induce longitudinal buckling, the claimants have 
found that inclining these at an angle is successful in 
cancelling the bending moment induced by the front 
end-load. Asymmetric side beads have also been 
used in vehicle body front-end longitudinals to con- 
trol crush characteristics in frontal impact. The beads 
act as fold initiators but are not fully effective as the 
comers of the members take greater load than the 
sides. In Fig 2, asymmetric comer divots are intro- 
duced as triggers and the arrangement particularly 
suits larger and heavier vehicles which have thick 
gauge longitudinals. Because the divots help to re- 
duce the bending moment of inertia at the cross 
sections where the beads are, this proposed design 
relies only on specially configured comer divots — 
with long and short axes arranged to pair with one 
another as seen here. Two-stage energy absorption is 
another approach, avoiding premature firing of the 
occupant-protection airbags . This is the object of the 
bumper configuration in Fig 3. Two-stage impact 
energy absorption is obtained by a soft outer bumper, 
supported on relatively weak collapse-tubes, de- 
forming under minor impact without firing the airbag. 
Heavier impact causes the more rigid inner bumper to 
collapse on its support tubes, with sufficient 
decelerative force to fire the airbag. Structurally 
efficient longitudinals are also needed behind the 
front-end. An improved structural efficiency for the 
platform underbody of an integral-bodied vehicle 
results from the use of a novel design of side-rail in 
Fig 4. The side-rail is a built-up beam having channel 
section flanges and a form of corrugated web in 
between. Preferably a pair of rails are sandwiched 
between upper and lower floor panels with the whole 
structure supporting a pillar-arch and cowl. 

Fig 1: Patented front-end by Fuji Jukogyo 

Fig 2: Patented 
controlled collapse 
system by Ford 

Fig 3: Two-stage 
energy absorption 
by Rover Group 




Fig 4: Patented longitudinal design by American 
Motors (GB 2 187686A) 

Fig 5: Patented body shell construction by Honda 
(GB2305639A ) 

Figs 6/7 : Patented side impact protection structure 
(and section) by Fuji Jukogyo (GB2306922A) 


Body shell integrity 

In giving structural integrity to the whole body shell 
frame a number of new approaches have been tried. 
In this body-shell made from rings, a light alloy 
integral framework for a car body is made possible. 
In the design of Fig 5, which affords minimum 
material wastage in the form of off-cuts. Almost the 
entire shell framework is made from extruded strip 
formed into polygonal looped rings. The rings are 
interconnected over the lengths of their side to obtain 
good structural integrity. 

There is much effort to concentrated on im- 
proving structural integrity in side-impact. Prevent- 
ing intrusion into the passenger survival space of the 
sill and rear quarter panel, during side impact, is the 
object of the design in Fig 6/7. At the intersection of 
sill and rear quarter panel, shown in the cross-section 
view, an additional longitudinal crash tube is intro- 
duced and the joint is braced across the vehicle by a 
U-sectioned transverse member which forms a box- 
tube when welded to the floor panel. 

Combining rigidity with energy-absorbency is 
a further requirement of the structural engineer when 
he considers the vehicle interior. The construction of 
vehicle body pillars and rails is generally such as to 
provide the highest possible strength and rigidity for 
a given weight of material and cost of fabrication. 
This conflicts with the requirement for a degree of 
energy absorption when these structural members are 
exposed to the vehicle occupants. So a better compro- 
mise is achieved by using the elements shown in this 
typical section, of a windscreen pillar, Fig 8. The 
conventional pillar section has an extra member 
sandwiched between its outer and inner ‘panels’ so 
that panel thicknesses can be reduced accordingly. 
Then the inner panel has an extra, U-shaped section 
spot-welded to its walls so as to provide an additional 
energy-absorbent panel of yet thinner gauge. The 
usual deformable plastic garnish panel then separates 
the absorber from direct contact with the occupant in 
an impact situation. Also shown in the sectional view 
are the door pillar and weatherseals which define the 
shape of the basic windscreen pillar section. 

Exploiting wall forces to grip occupants is 
another interesting approach. An unusually high 
degree of occupant protection in side impacts is 
afforded by the seat and airbag layout of Fig 9. The 
upper portions of the seat backrests are extended 
outwards to almost touch one another and the adja- 



cent areas of door trim. Platforms thus formed con- 
tain airbags which fire forwards but exert side pres- 
sure on the occupants by reacting against the door or 
sidewall trim. 

Extra protection from B -pillar impact can also 
be designed in. Means to permit the vehicle occupant 
to swing laterally in a pendulum effect, with pivot at 
the upper side rail, is proposed in the design of Fig 10. 
A yieldable portion of the B-pillar to sill joint col- 
lapses, with an adjacent section of the floor-pan, 
under side impact. The patent also covers proposals 
for controlled collapse of the B-pillar based on sepa- 
rate deformable portions. 

Taking the injury out of side impact is also 
possible with the arrangement seen in Fig 1 1. Nor- 
mally the B-pillar bends under side-impact at the 
discontinuity between the upper, window, and the 
lower, door, portions of the pillar, at occupant chest 
level. Here the centre-pillar has a designed-in strength 
discontinuity portion at seat level. In order to ensure 
this failure mode takes place, top and bottom anchor- 
age supports for the pillar also must be reinforced. 
This is achieved by using a box-sectioned pillar 
rooted to box-sectioned transverse members at roof, 
X, and floor level, Y, which make up a ring frame 
around the vehicle and include reinforcement of the 
floor centre tunnel, Z. 

An associated proposal. Fig 12/13, to that in 
Fig 6/7 comes from Toyota and covers the provision 
of an energy absorber within the trim of the door to 
further increase side-impact protection. Conventional 
cushioning devices added to the trim protrude into 
the passenger survival space. Here the cushions are 
provided within the thickness of the door and its 
associated trim case and configured to optimize the 
energy absorbency and allow the occupant to take the 
impact at hip level. 

Chassis/bodyshell elements 

Slotted construction for chassis frames is an interest- 
ing way of reducing the weight of vehicle structural 
elements. The possibility to make a lightweight 
vehicle chassis frame in aluminium alloy, without 
the need to form a complex one-piece tapered 
sidemember, is clear in the design of Fig 14. Rela- 
tively short lengths of constant section extruded 
sidemember portions are joined together by strong 
light interconnection members of hollow ribbed con- 
struction which can have such items as spring hanger 

Fig 9: Technique of occupant protection in side- 
impact by Mercedes-Benz (GB2309440A) 

Fig 10: Design for 
protection from the B- 
pillar by Fuji Jukogyo 
KK (GB2311259A) 

impact arises from this arrangement proposed by 
Fuji Jukogyo in Patent GB2292716. 

Fig 14: Patented sidemember construction by 
Dana in GB2313577A 



Fig 12/13: Associated proposal in GB2292913A, 

Fig 15: Technique for maintaining shear strength, 
patented in GB2203393A 

Fig 16: Oddments box doubling as structural 
member in patent by Daimler-Benz ( GB2312190A ) 

brackets integrated with them. 

Protecting chassis sidemember shear strength 
is a particular problem which occurs on skeletal 
semi-trailers, for container transporting, because of 
the comparatively shallow depth of the main longitu- 
dinal beams. This means that transverse, bolster, 
crossmembers of rectangular section cannot easily be 
inserted through them and welded in position. The 
cut-outs which pierce the sidemember webs can be 
too large to allow any residual shear strength in the 
sidemembers. In this arrangement by Crane Fruehauf, 
Fig 15, a pair of plates and connecting tube are 
inserted inside the crossmember so that the plates 
correspond to the cutouts in the sidemember. After 
welding up they replace the lost web area in the 

Even a dual purpose oddments-compartment 
can be used to enhance occupant protection in im- 
pact. The proposal for a centre-console locker in Fig 
16 includes provision for a measure of side-impact 
support and the possibility of having its contents 
warmed by heater ducting. The box structure has two 
cross bulkheads, two side ones, stiffened by folded 
flanges, and a floor bulkhead. The sides also form the 
outer walls of fore/aft running heater ducts. 

Stiffening up a convertible body is often neces- 
sary when the roof panel is removed from a car body 
shell. Increasing the torsional and bending rigidity of 
the rear-end of an open convertible car structure is the 
object of the design in Fig 17. A vertical structural 
wall forms the front-end of the hood stowage com- 
partment at the very rear of the vehicle. At its outer 

Fig 1 7: Patented method of strengthening 
convertible body, by Porsche (GB2308104A) 



edges, and in a cruciform across its leading side, rail 
structures react loads imposed on the body structure 
by the rear part of the chassis, the upper crossmember 
directing the loads to the flux point K. The forces 
reach this bulkhead via cranked box-section mem- 
bers, and a tapered pillar of substantial section, which 
connect to the sills and main chassis. 

A suspension-location structure can be used to 
supplement body shell structural rigidity. This ap- 
plies to Porsche in the chassis construction of the 91 1 
Carrera. Of particular interest is the rear suspension 
arrangement which involves a sophisticated sub- 
frame structure, Fig 18, which is elastically mounted 
to the body. The LSA lightweight strut axle com- 
prises a five-link wheel location system, the links 
being arranged in two planes, two forming the upper 
‘wishbone’ and three, including a tie-rod, forming 
the lower one. As the wishbones are effectively each 
a pair of tension/compression members, these are 
largely relieved of twisting or bending moments. 
Combined with its brakes and drive-shafts the whole 
is a pre-assembled module. Each suspension stmt is 
pivot-mounted at either end, to wheel-carrier and 
sub-frame respectively. The suspension ‘stmts’ are 
die-cast by the Acural process and are of C-section, 
with webs reinforced by diagonal ribs. These are in 
aluminium alloy AlSI7Mg with 240 N/mm 2 yield 
strength. The sub-frame members are of similar 
construction but use a H- rather than C-section. Total 
weight of the assembly is 17 kg. 

Many of the smaller body-builders consider the 
way ahead in structural efficiency lies with compos- 
ite bodies. Three dimensional weaving of glass-fibre 
preforms by the technique announced recently by 
Cambridge Consultants is likely to open up new 
opportunities for composite structures. By 
computeriszing the complex instmctions to a Jac- 
quard loom, which works by passing longitudinal 
fibres through loops in vertical wires, moved verti- 
cally as the shuttle carrying the thread moves be- 
tween them, automated construction of the preform is 
possible. A box beam having longitudinally oriented 
fibres in the flanges and at 45 degrees on the webs can 
readily be programmed. Vehicle seats, bumpers and 
side beams are seen as other applications as the 
technique allows curved forms and flanged edges to 
be achieved without difficulty. The main limitation 
of the process is the 50-55% fibre content limit in 
GRP needed for the resin to penetrate between the 

fibres in resin transfer moulding. A resin injection 
technology such as RTM or network injection mould- 
ing is seen as a key part of the automating process, for 
which production rates of between 10,000 and 100,000 
per year are seen as optimal by CCL. Fig 1 9 shows 
straight, curved and tapered parts which it is possible 
to construct, making the truck chassis frame to be a 
potential application as well as body side panels. It is 
even possible to envisage economical integral con- 
struction of monocoque van bodies provided a large- 
enough loom could be found to handle the size of the 

Structural/suspension innovations have also 
reached production in a Lotus model. Torsional 
stiffness of 6600 lbs ft / degree has been achieved for 
the open body structure of the Elan, with a steel and 

Fig 18: Porshe suspension sub-frame 

Fig 19: Parts made from the CCL process 



polymer composites combination structure, Fig 20. 
Body panels are produced by the VARI process. A 
constant curing temperature and newly modified 
resin systems have since improved the production 
cure time. Also a new preform system is used to 
produce the panels which enables sharper comers to 
be achieved. Body shape gives 0.34 drag coefficient 
and front/rear lift coefficients of 0.30 and 0.065 
respectively. The chassis structure has departed from 
the traditional Lotus pure backbone with the integra- 
tion of the composite platform which is a single-piece 
3 mm thick moulding, including inner sill, toeboard, 
heelboard, ‘A’ and ‘B’ posts. The resin is an isopthalic 
polyester reinforced by continuous filament glass 
fibre with additional reinforcement at the mating 
points with the backbone outriggers. The latter are 
made from 18 gauge steel which is E-coated and wax 
injected. Body panels are joined by elastomeric 
polyurethane adhesives. There is also a steel subframe 

inside rather than outside the front A-panel. 

A patented ‘interactive wishbone’ suspension 
assembly is mounted on a separate raft of heat-treated 
aluminium alloy. The raft allows very stiff fulcrum 
bushes to be employed but without passing road 
noise; it also permits abnormally low caster angle and 
accurately controls suspension geometry under all 
conditions. Torque steer effects are also reduced 
while allowing longitudinal wheel compliance with- 
out changes in handling. There is not the normal 
excessive loss of caster during braking, caused by 
deflection of the fulcrum bushes, and braking stabil- 
ity has actually been enhanced by offsetting the hub 
spindle rearwards relative to the steering axis. Track- 
ing stability is also optimized by commonizing cam- 
ber change geometry, with bump travel, on both front 
and rear suspensions. Roll centre heights (30 mm. 



front, 60 mm, rear) remain virtually constant for 
normal roll angles and steering geometry of 60 per 
cent Ackermann has been chosen. 

Many believe the achievement of optimum 
structural rigidity in car bodies will require the use of 
a punt-type structure with sandwich floor panels. A 
sandwich structure for a roadster chassis is seen in Fig 
21/22. Mazda engineers' released details of this 
sandwich panel sports-car project in SAE papers, in 
which the admirably simple ‘punt’ structure is de- 
scribed. This approach is felt to be justified in open- 
top cars which otherwise demand costly and weighty 
understructure modifications when derived from 
standard saloons. Structural efficiency was measured 
by calculating the strain-energy involvement of each 
panel by finite-element techniques. Aluminium alloy 
face and honeycomb core are used to obtain body 
bending rigidity of 1.0 X 10 3 kNm and torsional 
rigidity of 0.7 kNm/rad from the 100 kg . structure 
whose natural frequency in bending and torsion was 
above 40 Hz. The latter figure is well above ‘shake’ 
prone adapted saloon open-tops. Main floor panel is 
100 mm thick and attaches to the structural side sills 
as shown inset. 

vehicle impact. Inflation pressure is also controlled 
so as not to over stress the child during this event. A 
further airbag is mounted in a play-table which is an 
integral part of the seat. A weight sensor and infra-red 
sensor determine the mass and size of the child and a 
preferred arrangement also includes inflatable shoul- 
der and lap belts to increase the restraint effect. The 
airbag and gas generator assemblies are preferably of 
the plug-in type, the base of the sockets being elec- 
trically wired to the main harness of the vehicle, as 

Car body systems 

Fig 23: Child safety seat by VW in GB2290505A 

Occupant restraint 

Occupant protection has also been considerably af- 
fected by developments in vehicle seating and sur- 

Preventing submarining or similar motion of a 
child in its safety seat is proposed in this Volkswagen 
patent, Fig 23. The shell of the seat is a carrier body 
for airbags which inhibit motion of the child during 

Fig 24: Patented pre- 
tensioner by Allied- 
Signal (GB2304027A) 



Fig 25: Safe install pretensioner by Allied Signal 
SpA GB2315983A 

Fig 26: Energy-absorbing anchorage by Tokai 
Rika Denki Seisakusho (GB2307168A) 

Fig 27: Side airbag within seat back by Allied 
Signal (GB2305638A) 

Pretensioners for seat belts are another safety 
system receiving attention. In Fig 24, a relatively 
easily assembled and installed seat belt pretensioner 
mechanism has a strong piston locking element which, 
while allowing free movement of the piston in the 
webbing tensioning direction, locks the piston in the 
web-loosening direction. The cylinder is internally 
threaded at one end for setting the pretension and at 
the other has a toothed profile into which the locking 
element engages as well as two O-ring seals spaced 
at one tooth pitch. 

The hazard of handling seat-belt pretensioners, 
while armed is overcome in the design of Fig 25, 
which also makes it difficult for unauthorized dealers 
to fit pre-used units and provides a visual indication 
to safety inspectors when this happens. The action of 
fitting the unit to the vehicle causes the lever of the 
arming bracket to be actuated and the bracket incor- 
porates a flange for limiting access to the fixing 
screws. The arming bracket is secured to the retractor 
frame by a twin headed arming screw, the heads 
being separated by a weak portion of the stem. On 
tightening, the stem shears to detach the outer head 
and leaves the inner one protected by an enveloping 
sealing ring. 

Energy-absorbing belt anchorages are another 
approach to lessening impact injury. In Fig 26, an 
upper seat-belt anchorage is provided which incorpo- 
rates an additional measure of energy absorption. 
The body of the anchorage which supports the web- 
bing has a specially shaped support-pin passing 
through it, and the adjacent body centre-pillar, which 
engages an exterior reinforcement plate in a special 
way. The energy absorber is positioned around the 
circumference of the outer end of the pin and is 
arranged to expand outwards under crash loading on 
the interior belt mounting. 

Out-of-view side-impact airbags protect pas- 
sengers from impacts in other directions. In the 
design of Fig 27, no obtruding into cabin space, or 
detraction from the interior styling, results from a 
modular side-impact protection airbag which fits 
within the seat back. A specially constructed ex- 
truded seat-back support housing has a fold over part, 
behind which the bag (plus inflator) is stowed, which 
allows the bag to deploy forwards, between the 
occupant and the door. 

In another ingenious design the frontal airbag 
module is the horn push. Fig 28. Provision of tactile 



indication of when sufficient pressure has been ap- 
plied to sound the horn is inherent in this steering 
wheel configuration which also prevents unwanted 
application of the airbag. The airbag module is 
movable with respect to the rim of the wheel and is 
mounted on three resilient elements which are also 
electrical contacts of the horn switch. The force/ 
deflection characteristics of these elements are such 
that the resistive force rises to a maximum during 
initial movement and then falls to a minimum and 
rises once more to a second maximum with continued 

Side protection is also possible with an open 
window in the design of Fig 29. Protection of vehicle 
occupants during side impact, when a side-window is 
either broken or open, is the objective. A strap that 
threads through the deflated airbag is anchored to 
structural pillars either side of the window. When the 
bag inflates, the strap is tensioned by the disposition 
of inflating cells around the strap such that the 
occupant is constrained inside the window opening. 

One school of thought believes that occupant 
protection is best served by having seat belts an- 
chored into the seat, particularly in open roadsters. 
Seats on the M-B 300 SL, Fig 30, have ten way 
electric adjustment with memory pre-set and a struc- 
ture which includes anchorage for the shoulder belt 
inertia reel. It was designed with the aid of the 
material supplier, Norsk Hydro. Possibility to adjust 
the seat-belt harness to almost any occupant size is an 
important contributor to safety in an open car which 
otherwise has to rely on safety-reel anchorage by the 
half-height C-pillar. Since the crash loads on the seat 
from the safety harness are very considerable the 
integrity of the structure and its slide mechanism is 
crucial and required a 5800 element FE analysis in 
the SL design 

A particularly compact seat-belt pretensioner 
has been patented by an Asian manufacturer. Avoid- 
ance of a long cylinder and rack to effect pre- 
tensioning of a seat-belt retractor is one of the objec- 
tives of the compact design in Fig 31. The retractor 
uses gas pressure to move the rack and rotate the 
pinion gear. A clutch mechanism transmits motion 
between the member attached to the pinion gear and 
a belt take-up spindle, the clutch mechanism being 
driven through step-up epicyclic gearing. 

More effective seat-belt locking is claimed for 
another ingenious design. In particular, a faster, more 

Fig 28: Airbag module as horn push by Autoliv 
( GB2309123A ) 

Fig 29: Protection from side-window shattering by 
Autoliv Development (GB2312877A) 



Fig 31: Compact pretensioner by NSK 

Fig 33: Adjustment mechanism by Dunlop Cox, 

effective, release of the clamping wedge and spool 
assembly of a seat-belt retractor is achieved, in the 
design of Fig 32, when the loading on the belt drops 
below the spring bias force. The spool assembly is 
mounted so that linear movement with respect to the 
wedge housing pushes one or more clamping wedges 
vertically to lock the seat belt. Ideally, the spool is 
coupled to the wedges by coil springs and it is 
supported in a slider which is also the spool-assembly 
frame. The slider moves linearly with the spool to 
activate the wedges. 

An improved manual height/tilt adjuster for a 
seat has solved a long-standing problem. Seat adjust- 
ment mechanisms using double pinned discs to en- 
gage notched racks have previously been impractical 
because of the difficulty of turning a coaxial 
handwheel through 1 80 degrees. For the design in Fig 
33, a 3-pinned disc allows a more manageable angle 
for hand-turning and incorporation of such an actua- 
tor in the mechanism shown, and when fitted either 
end of a vehicle seat slide, allows a convenient way 
to provide front and rear height adjustment. 

A simplified headrest adjuster is also a wel- 
come advance. In many small cars the headrest has 
little use, apart from the protection afforded against 
whiplash injury, as different sized drivers cannot 
adjust the fore-aft position even to make contact with 
the rest. In the design of Fig 34, lighter weight 
construction and improved comfort follow from a 
locking mechanism design for a seat head support 
which allows an adjustable position for the rest. Key 
to the mechanism is an actuating lever within the 
retaining yoke of the headrest. At its top end is a 
release button and at the bottom end the locking 

Fig 32: Improved 

belt locking by Allied-Signal (GB2300797A) 

Fig 34: Simplified head rest adjuster by 
Volkswagen ( GB2302 706A ) 



Controlling headrest position is also improved 
by a mechanism for helping to prevent whiplash 
injury. The design in Fig 35 incorporates an antici- 
patory crash sensor which signals when an external 
object at the rear of the vehicle has greater than a 
threshold closing speed. Another sensor determines 
head location with respect to the headrest. A proces- 
sor works out the correct headrest repositioning to 
minimize injury and actuates the mechanism accord- 

A simpler seat slide control that would seem to 
favour powered drive is seen in Fig 36. Substantial 
simplification of a seat slide and adjuster mechanism 
follows from the adoption of a system based on belts 
and pulleys rather than a serrated rack and pawl. Four 
seat-mounted pulleys are involved plus a simple 
brake mechanism, at the leading edge of the seat, 
which clamps the belt in any position and therefore 
provides finer adjustment than is possible with a 
toothed channel. 

A quieter belt-retractor for a seat belt is seen in 
the interesting design of Fig 37. It enables the 
ratcheting noise to be eliminated, of the pawl drag- 
ging over the locking teeth of a seat belt retractor 
mechanism, on rewind. A spool with a ratchet is used 
for carrying belt webbing while a rewinder is used for 
rotating the spool. An upturned cup-shaped member 
acts as a sensor for detecting acceleration of the 
vehicle over a given threshold, causing a lever to 
move upwards and move a pivotally mounted pawl to 
lock the spool. A blocking arm, frictionally coupled 
to the spool, disables the pawl when the spool is 
rewinding, its movement being limited by arcuately 
spaced abutments. 

Doors, windows and panels 

An interlocked split-tailgate is a way of overcoming 
difficulties that have arisen with two-piece tailgates 
on cross-country vehicles and estate cars — of inde- 
pendent securing of the upper (window) portion and 
the lower (flap) portion when long protruding loads 
are required to be carried. In the design of Fig 38, a 
gas strut is used to automatically open the window 
portion when the latch is released. However, if the 
window is open when the flap is open, it cannot be 
closed unless the flap is closed first, whether the flap 
is hinged horizontally or vertically. The flap can also 
be independently locked to the body via the D-pillar. 
In the closed position the window rests against a seal 

Fig 35: Controlling headrest position by 
Automotive Technologies (GB2301906A) 

Fig 36: Patented design for a power-slide by 
Rover Group ( GB2306880A ) 

Fig 37: Quieter belt-retractor by Allied-Signal 

Fig 40: Easy-assemble 

Fig 39: Door seal by door strip by Draftex 
Draftex (GB230994A) (GB2293191A) 



(a) which extends around to the roof, via the side 
pillars. In its lower part the window rests against seal 

(b) on the flap and the window is locked to the flap 
with a mechanism that prevents the flap being opened 
while the window is closed. The window raising 
linkage ensures no damage can be done to it when the 
flap is opened and closed. Reference to the inset 
enlargement of the window hinge linkage shows the 
window to have a horizontally hinged tilting lever 

(c) , with a roller attached to it, which has a second 
lever (d) pivoted to it, the angular motion of which is 

Fig 38: Opel split-tailgate proposal covered by 
Patent GB 2290060A 

Fig 41: Winder mechanism by Nifco Inc 

limited by a stop (e). If the unlocking mechanism is 
actuated when the window is closed, the window 
moves from the closed position into the partially 
open position (f). Further automatic opening of the 
window can then only occur after a manual push 
because the ‘power’ lever arm is then too short. 
Closing of the window under manual pressure is 
arranged so that the window pivots a given distance 
until the roller rests again on its track. Under further 
pressure, the tilting lever pivots about axis (g) and the 
roller moves on its track until the shackle catches 
behind the latch at the bottom of the window. 

Improvements in door sealing are the subject of 
many patent applications. In the design shown in Fig 
39 the upper rail of the door frame is made with a 
lipped channel section whose open-end faces in- 
wards to receive the specially designed weatherstrip 
section. The lip of the channel holds the weatherstrip 
in place after its integral locking contour has snapped 
into position. The door periphery is then sealed, on 
closure, in two places by direct compression of an 
inbuilt ‘balloon’ and by shear deformation of the 
outer edge of the seal. 

Weatherstripping without tools is possible with 
the design in Fig 40. The ability to assemble and 
remove a door weatherstrip without special tooling 
follows from the metal reinforced elastomeric strip 
fitting over flange A in the usual way but being 
locked into position by a snap-over section of the 
resilient strip B which can be withdrawn with modest 

A successful yet inexpensive window winder 
for curved glass taxed designers for many years. 
Unless a complex compensating mechanism is intro- 
duced into a window regulator used for curved win- 
dows, there is a tendency for a wrenching force on the 
glass-holding clamp. This is because the clamp rises 
in a linear path while the glass rises in a curved one. 
In the design of Fig 41, the glass clamp is rotatably 
mounted to the raising linkage so no unwanted torque 
is induced. 

A long-life window-winder drive is promised 
by the design of Fig 42. Deterioration in operation of 
window regulators can occur in conventional chain 
driven systems due to stretch of the associated wire 
rope — and the inability of the spring-tensioner to 
compensate for the stretch. Here an integrally formed 
chain has interior teeth and a flat outer side with a 
wire rope wrapped between inner and outer sides. An 



electric drive is used with the endless chain, which 
engages driver and driven gears enclosed by a chain 
protector. The chain-runs lie in flexible tubes. 

Easy access for door hardware assembly is the 
object of the design in Fig 43. A further improvement 
in modular door construction comes in this design in 
which hardware mechanisms such as door latch/lock 
and window regulator are built into a central frame- 
work. It also forms the structural integrity of the door 
and supports items such as rear-view wing mirrors 
and hinges. Inner and outer non-structural skin and 
trim panels are snap fitted to the central framework 
after it has been fully assembled with the necessary 
hardware, without access hindrance. 

An extending door handle to ease operation is 
seen in the design of Fig 44. The problem encoun- 
tered by relatively small stature people, in driving 
two-door cars, of being able to reach the door handle 
to close the door from its fully-open position, is 
addressed by this design. A strap-type door handle is 
made extensible by means of a linkage between the 
door and its frame-surround. As the door is pivoted 
open, one end of the resiliently flexible handle is 
moved towards the other end causing it to bow 
outwards and be easier to grasp by the vehicle occu- 

The current sophistication in door latching and 
locking systems is exemplified by the Bosch inte- 
grated door latch, lock and auto-closing system. The 
company have revealed details of their electronically 
controlled automatic door closing and security sys- 
tem in a recent paper 2 . Only one actuator motor is 
involved and the functions of the device are inte- 
grated into one system. In an effort to minimize the 
number of mechanical parts, for closing and locking, 
only motor, gearbox, driving/stop plates, guide mecha- 
nism and ratchet are required; door slider and lock 
slider are located on one bracket. Door position is 
seen by the displacement of the drop latch, monitored 
by two proximity switches, while that of the driving 
plate is seen by two Hall-effect sensors. Shown at Fig 
45(a) are the door in open and the first ratchet 
positions; the auto-closing system is inoperative when 
the closing pin is introduced in the first ratchet 
position as this actuates the controller via a proximity 
switch. The ratchet then holds the drop latch and 
prevents the door from springing open. The driving 
plate is connected to the gearbox output which turns 
the drop-latch into the closed position via a sprung 

Fig 42: Long-life winder by Chin-Yun Huang 

Fig 43: Easy access to door hardware by Honda 
(GB2315513A ) 

Fig 44: Extending handle by Prince Corporation 



Fig 46: Rack and pinion mechanism by Rover 
Group, covered in Patent GB2291 109 A 

stop-plate. The drive plate then completes one revolu- 
tion and the door can then be mechanically opened by 
the exterior/interior handles. A console switch signals 
the actuator, via the CAN bus, to enter the central 
locking mode. The driving plate is then motored to the 
door-locked position (b). The driving pin on the stop 
plate prevents the drop latch from being handle- 
opened from the outside — but interior handles will 
turn the stop plate against its spring with the slider. 
Turning the locking key signals the actuator to enter 
the anti-theft mode in which the driving plate takes up 
the position monitored by the sensor; both inner and 
outer handles are then disconnected as in (c). 

Another way of foiling the thief is with a 
vehicle locking system which overcomes the grave 
limitation of many existing systems, that once the 
thief has prised the lock from the body the latches 
can be operated by the pull rods linking them to the 
lock. In this arrangement, Fig 46 , the lock actuator 
is a key-driven rack and pinion, the rack of which is 
connected to a Bowden cable that links with 
the latch. The lock, rack and pinion, and the end of 
the Bowden cable are all located in a casing such that 
the locking function cannot be operated without the 
key to the lock, even if the lock is prised out of the 

Electrical Door Closing 
Door open 

Door In first Ratchet 

Anti-theft Mechanism 

Fig 45: Integrated latch/lock system by Bosch 



A way of minimizing operating forces on a 
central-locking system is the design of Fig 47 in 
which the electric motor only needs to be sized 
according to the operating forces of the locking lever 
system. Manual operation can also be carried out 
only against the resistance of these forces, in the 
event of motor power failure. Thus a rotary latch, 
pawl and release lever have an operating lever which 
comprises at least one operating lever plus a locking 
lever system . This involves an interior locking lever, 
central locking drive and locking element connected 
to it. The drive is a reversible motor with power take 
off having an eccentric control pin. The pin is con- 
trolled in a way that it rotates in either direction to 
move the central locking element into ‘locked’ or 
‘unlocked’ position. This element has a forked re- 
ceiver with control surfaces on its sides which cause 
the interior locking lever to be rigidly connected to it. 
Part of the control pin arc of rotation is outside the 
forked receiver and the central locking element has 
a stop face on both sides, for the control pin. Control 
movements of the pin are restricted by its striking one 
of these faces and the motor being switched off as this 

A particularly compact power door latch is 
obtained from the design in Fig 48. Existing pow- 
ered-closing devices for vehicle doors cover a large 
projected area of the door. This is because the cancel- 
ling lever normally projects beyond the output mem- 
ber. In this design a compact layout is achieved by the 
cancelling lever overlapping the output member, as 
well as the member connecting the output member to 
the intermediate lever being shorter. The output 
member is rotated by a motor, moving the latch from 
a half-latched to a full-latched position. The cancel- 
ling cam surface is brought into contact with the 
connecting member to disconnect the output mem- 
ber, and the intermediate lever, when the opening 
handle is turned. 

Another body mechanism which receives 
considerable design attention is that for convertible 
hoods. A compact folding-roof arrangement for boot- 
stowage is covered in the design of Fig 49, which 
maximiszes the space within the stowage cavity. To 
pivot the boot-lid frame upwards a hydraulic jack is 
extended and a slot in the lower end of the jack slides 
down over a fixed pivot. This initial motion of the slot 
causes the foot of the jack to pull on a cable, to unlock 
the boot lid. Further jack-extension raises the boot 

Fig 47: Reduced locking force system by Kiekert 
AG (GB2292768A) 



Fig 48: Compact door latch by Mitsui KKK Kaisha 
( GB2313408A ) 

Fig 49: Boot-stowage of folding roof by Mercedes- 
Benz, covered in Patent GB2300671A 





lid; jacks are positioned at each side of the boot, 
behind the wheel arches. Leaf springs are positioned 
at the side of the slots to stabilize the pivot positions. 

Fig 51: Sliding door drive by Mitsui 

Fig 52: Reliable bonnet latch by Bloxwich 
Engineering (GB2312921A) 

Trim and fittings 

A mirror system without rain-blur is seen in the 
design of Fig 50. Substitutes for conventional rear- 
view mirrors have been proposed which avoid the 
considerable protrusion of the mirror from the sidewall 
of the vehicle. However, a drawback of such systems 
is the blurred image obtained in wet weather when the 
image is distorted by rain droplets. In this patented 
design, a housing on the vehicle side has a concave 
mirror inside it to reflect light collected by a circular 
concave lens at the rear of the housing. The lens is 
coated with a water repellent and is mounted in 
bearings so that it can be spun by gear teeth, on its 
periphery, driven by a motor pinion. 

An improved sliding-door drive which does 
not require exclusive use of power is suggested in Fig 
5 1 . A clutch mechanism, for transmitting power from 
motor to movable member, has racks fixed to station- 
ary members and engages with gears to drive the 
movable member. The clutch unit is supported on a 
swinging member, spring-biassed to a central neutral 
position. Gear A is engaged with the output gear B, 
and the stationary rack, while gear C is frictionally 
coupled to gear A. On energizing the motor the 
assembly swings away from the neutral position to 
engage with the movable-door drive gear. 

A more positive and reliable bonnet release 
mechanism is proposed in the design of Fig 52. This 
improves on the conventional type comprising a 
headed striker pin mounted on the bonnet and adapted, 
as the bonnet is closed, to pass through an opening in 
a catch plate attached to the grille surround and 
become engaged by a catch lever. Here a latch for 
retaining a striker pin does so by the action of a 
latching member in the latch on the head portion of 
the pin. The latch has a resilient member, which is 
chordal to the bore, to retain the striker pin in the 
latch. A release ring rotates about the axis of the bore 
and has an abutment which engages with the resilient 
member to spring it from the bore and release the 
strike pin. 

Cushioning a bonnet top impact is the object of 
the design in Fig 53 for a pedestrian protection 
device. In the cases of frontal impact with the vehicle, 
uses an external airbag triggered by a deceleration 
sensor. Protection is given particularly against vio- 
lent head impact from the whiplash effect of the body 
folding after leg impact from a low bonnet. The 
airbag is stored in a protective cover which ruptures 



on inflation of the bag, the cover being contained in 
a nudge-bar extension of the normal bumper bar, 
which is itself covered with soft material to minimise 
leg injury. 

Protection of the occupants is the object of this 
controlled-collapse trim system in Fig 54. Providing 
a crash impact waveform with a number of progres- 
sive deceleration peaks is an objective of a trim 
finisher for fitting to the interior of the windscreen 
pillar. Earlier attempts to do this have usually re- 
sulted in undue bulk of the trim member with conse- 
quent intrusion into the passenger space and/or ob- 
struction of the front doorway. Here the generally 
channel-section plastic trim member has reinforcing 
ribs at right angles to its surface and arranged with 
different clearance dimensions from the screen pil- 
lar. Fastening of the trim to the pillar is via integral 
bosses parallel to the ribs which engage with a sprung 
plate with lugs engaging spring-clip fasteners fitted 
in holes in the pillar. 

A simplified level-sensing system for lamp 
adjustment is seen in Fig 55. Undue complexity, and 
a large number of individual parts, in existing vehicle 
level-detection devices have resulted in costly solu- 
tions for automatic on-vehicle headlamp adjustment 
and suspension levelling. In this proposal a sensor is 
used having just two angularly moveable parts, re- 
spectively coupled to the vehicle body and suspen- 
sion. The ‘fixed’ part is attached to a body-structure 
member in a way that allows the sensor to be coaxial 
with the fulcrum of a wheel suspension control arm, 
an attaching leaf-spring from which holds the ‘mov- 
ing’ part of the sensor. The rubber bush of the 
suspension control arm compensates for any motion 
of the arm at right angles to the axis of rotation of the 
sensor moving-part. 

A system of fibre-optic light division for head- 
lamps is seen in the design of Fig 56. This headlamp 
design uses fibre-optics so that only one high inten- 
sity discharge bulb is required for both main- and 
dipped-beam operation. Light conducted from the 
upper (main beam) source, by the optical fibre at the 
rear of the lamp, serves as a source for the lower, 
dipped-beam unit, a shutter mechanism controlling 
the amount of light passed. The arrangement is also 
said to improve colour distribution between each 
source, give better balance of luminous intensity and 
thus better forward vision. 

Body system innovation in the fuel tank region 

Fig 53: Pedestrian protection device patented by 
Concept Mouldings Ltd (GB2316371A) 


Fig 55: Simplified level-sensing by Mercedes-Benz 



includes the fuel delivery system of Fig 57. Simpli- 
fied assembly and test can take place of this fuel 
distributor system, which is an add-on unit to a fuel 
tank, rather than fabricated within it. The carrier 
casing, which is mounted to the fuel tank, has a non- 
return valve, bottom right, and an ingenious jet-pump 
arrangement, for filling. The jet-pump, left side of 
carrier, involves excess fuel flowing back through a 
duct from the engine. 

An improved fuel tank, overall, is covered in 
the patented design of Fig 58. Prevention of overfill- 
ing on refuelling, and avoidance of any bad effects of 
vapour formation on vehicle acceleration, are the 
objectives of this fuel tank design. Two separate 
ullage spaces are formed by means of an inverted 
canister suspended from the roof of the tank, which 

Fig 56: Fibre-optic light division system by Fuji 
Jukogyo (GB2315538A) 

forms a vapour reservoir. An acceleration-sensitive 
valve controls the opening between the two ullage 
spaces which are also interconnected by a small bleed 

An exhaust clamp without bolts is an interest- 
ing piece of underbody innovation, Fig 59. Problems 
of looseness in an exhaust pipe assembly, following 
the fixing clamp nuts becoming untightened in serv- 
ice, are overcome in this relatively easy-to-assemble 
design. The ubiquitous bolted exhaust pipe clamp is 
replaced by a spring-clip-shaped member which at- 
taches to a support arm by notched engagement with 
its flattened end. The rod is hung from a flexible 
mount in the usual way. 

and weight saving 

Attention to underbody panel shape is also important 
in extracting the last mph of speed. Dr Dominey of 
Durham University 3 points out that the stability of 
regulations has led to a convergence of design which 
means that competitive circuit performance now 
depends on fine tuning of aerodynamic design. As 
ground clearance beneath the front wing is reduced 
flow can no longer tolerate the pressure gradient and 
the aerofoil stalls in the limiting condition. The 
limited permitted span of the aerofoils makes it 
necessary to control the associated tip effects. The 
end-plate solution aimed at maintaining approximate 
two-dimensional flow increases loading at the wing 
tips but does not eliminate strong tip vortices — for 
which semi-tubular guides along the lower edges of 
the end-plates has been one of the ways of containing 
them. To overcome the compromising effect of the 
front wing on engine cooling performance, the use of 
flapped aerofoils has been introduced — with chord 
and camber at the outer end of the flap greater than at 
the chassis end. Combining this with taller and 
narrower engine air-intakes ensures that only the 
wake from the inner section of the wing impinges on 
the intake. With rear-wing aerofoils the requirement 
for high downforce to be generated from already 
disturbed airflow has necessitated the use of highly 
cambered multi-element configurations which are 
inefficient in terms of drag. The trailing edge of the 
wing is governed by regulations while leading edge 
position is chosen as a compromise between a high 
one better exploiting airflow and a low one allowing 




a larger aerofoil to be used. The undertray of the car ent sections of the floor. Also obtained were interest- 

is the other important factor, Fig 60, air flowing ing results due to the progressive alteration of rear- 

beneath the car leading to a high-pressure stagnation end shape between the limits of notch- and hatch- (3) 

point on the upper surface and a separation bubble and of adding on and progressively increasing the 

below the entry lip. Addition of a diffuser creates a depth of a front spoiler (6). 

further zone of low pressure and its design is crucial An innovative body construction material is 

to obtaining marginal advantage. low density structural RIM. An LD-SRIM system has 

How the influence of a vehicle underbody been trailed as a RRIM replacement for production of 

affects aerodynamic drag was also discussed by automotive interior trim panels via the Strappazini 

Rover engineer J.P. Howell 4 . He showed, Fig 61a, at process. Trials at United Technologies Automotive 

(a) the contributions made by different elements of confirm that ICI’s MDI-based polyurethane LD- 

the body on the overall aerodynamic drag of a saloon SRIM system has exhibited advantages over the 

car. By fairing the complete underbody of an 800 commercial RRIM system. The Strappazini process 

saloon it was found possible to improve drag by is a patented method of moulding in place multiple 

factors of 0.39 and 0.46 for a notchback and hatch- covers with well finished joints and seams. The 

back configuration respectively. So as to gauge the process is useful in moulding complex automotive 

effect of progressively adding fairing panels (b), interior trim components, including interior door 

wind-tunnel tests were made with successive in- panels. ‘RIMline’ SL-87339, a system containing an 

crease in the numbered panels to obtain the results in internal mould release (IMR), was tested in a Cor- 

(c). The extremities of the floor gave the largest gains vette RRIM door panel tool at UTA. In a closed 

but there is evidence of interactions between differ- mould, with only slight variations in mould tempera- 

tures and throughput of the current production equip- 
ment, the ‘RIMline’ system was poured directly onto 
a glass mat using the Strappazini process. Multiple 
releases were demonstrated with ICI’s new IMR 
technology. During initial evaluation, 15 releases 
were achieved. Based on previous production expe- 
rience, it is expected that in a plant environment at 
least 30 to 40 additional releases between external 
mould release applications can be achieved. In com- 
parison, the current RRIM system is not available 

Fig 62: GM Corvette doors: front in RRIM rear 
view in LD-SRIM 

with IMR technology. 

The LD-SRIM part. Fig 61b, weighed 20 per- 
cent less than the same door part moulded from the 
RRIM material. This was accomplished without op- 
timization of part thickness or density. An additional 
20 to 25 percent weight reduction can be achieved by 
reducing part thickness due to the superior strength 
properties of low density SRIM composite. RRIM 
requires substantially thicker sections to meet stiff- 
ness and thermal stability requirements as well as 
adequate impact resistance. 

1. Front 
2- Sides/Roof 

3. Rear 

4. V^eels/VJheel arches 

5- Underfloor 

6- Ocoling 

Fig 61: Drag factors involved 



CV systems 

CV chassis-cab configuration 

Advanced-concept trucks are the trendsetters in truck 
design; DAF's future concept vehicle (FCV), Fig 62, 
is a fully driveable working unit whose features all 
have prospects, individually, for eventual production 
fitment. The chassis-frame is a fabricated rectangular 

design gives high torsional rigidity necessary to 
mount the independent suspension assemblies. Out- 
board of it are also hung fuel tanks and air reservoirs, 
while inside houses the major elements of the 
drivetrain. The apertures in the box tube are for 
access to various parts of the drive train, the front 
rectangular one giving engine access. A flat-bottom 
sump is fitted to the engine which is installed as low 
as possible within the box -tube, forward of the front 
wheels centre-line. 

While engine, semiautomatic transmission and 
prop-shaft are otherwise conventional, the rear wheel 
drive is novel, in the form of a ‘double drive differ- 
ential system,. A torque apportioning device trans- 
mits two-thirds of it through a straight-through shaft 
to the second axle, the other third passing, via an 
annular gear drive, to the input pinion of the lead 
axle. Being a hypoid gear, its pinion offset allows the 
through-shaft to pass beneath the cross-shaft of the 
front differential. A short propshaft then connects the 
drive to the second axle. Independent suspension is 
fitted to both front and second ‘axles’, the latter both 
driving and steering. Special low-profile air-springs 
are employed which bear on saddles on the stub- 
axles. Lower links of the wishbone suspension, whose 
roll-centre is at ground level, are cranked downwards 
to clear the inboard disc brakes. In the second axle, 
drive extends outwards to the wheel hubs via a ring- 
type kingpin housing. Suspension travel is 80 mm to 
bump and 160 mm to rebound stops. For the third 
axle, twinned wheels on special two-flange hubs are 
supported by stub axles on a rectangular frame. Its 




Fig 62a: DAF concept 
truck, with suspension and 
drive detail above 

Fig 62b: Concept truck suspension and drive detail 

slender, deep sidemembers have brackets extending 
outwards, beyond the inner road-wheels, to give as 
wide a spring-base as possible. 

Concept has become production in the case of 
the important breakthrough which has been made by 
Dennis Specialist Vehicles. It involves a dedicated 
space-frame chassis for the new Rapier fire-appli- 
ance vehicle, Fig 63. Dramatic improvement in ride 
and handling is claimed over the conventional lad- 
der-frame and leaf-spring layout — with the avail- 
ability of a more rigid frame to mount a suspension of 
greater sophistication. Over a tonne of weight has 
been pared from the conventional structure and a low 
deck height has also been achieved — with a particu- 
larly low-slung engine. The latter has permitted the 
cab to be mounted 250 mm lower than usual, remov- 

lit! CO;! 


Fig 63: Dennis Rapier space-frame chassis 

ing the need for entry steps. The entire crew cab tilts 
in one piece and the lightweight body is made in 
corrugated aluminium alloy sheet over an extruded 
section frame — incorporating water tanks made 
from plastic. Substantial coil springs suspend both 
front and rear wheels and, at the rear, wheels are 
independently suspended by a double wishbone sys- 
tem. Body roll, under 0.6 g lateral acceleration, has 
been reduced to 3.1 degrees compared with 8.5 
degrees for a comparable conventional vehicle. 

Cab/body fittings 

Recent patent specifications reveal a number of the 
detailed improvements which have been made to 
truck systems. An example is the suspended cab-tilt 
pivot design in Fig 64. Here a truck cab suspension 
which is sympathetic to the cab-tilt system, without 
causing discomfort to the occupants, is proposed. 
The cab itself hinges on the uppermost pivot of the 

Fig 64: Soft-mounted cab tilt by ERF, in Patent 
GB2 188884 A 



0 h 

Fig 65: Return-load tanker by Clayton 
Commercials ( GB2298830 A ) 


Fig 66: Variable cube body by York Trailers in 
Patent GB2198091A 

Fig 67: Air-bag system vari-cube system by JR 
Cramp , covered in Parent GB229931 1A 

mechanism shown while beneath it a second pivot 
forms the scissor linkage of the suspension mecha- 
nism — based on a rubber ‘doughnut’ compression 
spring and hydraulic damper. 

A return-load tanker is the subject of the inter- 
esting patented design in Fig 65. The classic diffi- 
culty is tackled here in the operation of tankers and 
bulkers, not being able to fill the tanks with a return 
load, which might contaminate the principal product 
carried. According to the claimants, existing efforts 
to build dual purpose vehicles, carrying both dry 
goods and flow products on flat-topped tankers, have 
involved overweight structures. Here the lower por- 
tion of the vehicle, for flow products, comprises 
typically two large cylindrical vessels which form 
substantial structural members of the vehicle. Cra- 
dles around the vessels support the upper floor for 
carrying dry goods as well as proving a mounting for 
the running gear of the vehicle. 

A number of attempts have been made to 
produce ‘variable cube’ bodies. A means of, literally, 
raising the roof of this van body is proposed by the 
design in Fig 66. An air-inflatable bag is positioned 
between the body wall and a lever which compresses 
the bag. Inflation of the bag moves the lever and pulls 
cables to operate the roof-raising device. Shown in 
this example mounted to the step of a step-frame 
trailer, cables from the pulleys shown pass beneath 
the chassis to operate jacks in the body comer posts. 
Another proposal also concerns raising the roof with 
airbags. Previous efforts have made been made to lift 
the roof of curtain-sided vehicle by an amount corre- 
sponding to the depth of the projecting border which 
covers the curtain tracks, to allow easier loading of 
tall cargo items. Because such efforts have thus far 
involved cumbersome mechanisms, the design of Fig 
67 is aimed to simplify the approach by the use of air- 
bags. A scissor linkage is associated with the air-bag, 
positioned at the top front/rear comers of the vehicle 
body. The compressed air supply would normally be 
reservoir tanks fed by the vehicle’s air-compressor 
and air-lines would ran through the roof support 

Improved cargo retention is the object of the 
design in Fig 68. Problems of complexity, and asso- 
ciated proneness to wear, of existing cargo retention 
devices for goods vehicles can be reduced by the 
adoption of this mechanism in the design shown. 
Channels on each side of the deck anchor bracing 



beams, by means of a socket beneath each channel. 
The beams are either stowed within the channels or 
tilted upwards to brace the load. The beam is released 
from the load-retention position by raising it towards 
the vertical and then displacing it axially, from where 
it is pivoted to the stowed position. The interior of the 
beam encloses a restraining strap which is withdrawn 
to secure the cargo, the strap retracting into the beam 
automatically when not in use. 

The object of the Fig 69 design is to provide an 
easy-access curtainsider. Means to avoid complex 
slider mechanisms for the body-side support pillars 
of a curtainsider are provided in the arrangement 
shown. In order to move the pillars to ease side- 
access of bulky loads, the pillars are mounted on a 
hinged linkage. 

Hinged comer-posts for a curtainsider are cov- 
ered by the design of Fig 70. The need to provide 
front-comer posts, of curtain sided CV bodies, which 
are faired in plan view to achieve aerodynamic 
efficiency, results in a dead space behind the section 
profile when loading the container body by fork track 
from the side. In order to avoid a secondary loading 
operation, to enable pallets to fill the dead space, 
hinged post sections are suggested. When the flaps 
are in their closed position the side-curtain is wrapped 
around them and secured in socket catches prior to 

Considerably reduced effort to open and close 
the side-curtains of a curtainsider track-body is 
claimed in the interesting roller arrangement of Fig 
71. The roller track section is profiled to allow tilting 
of the rollers as they roll along the track. The rollers 
comprise ball bearings with convex sectioned tyres 
of hard plastic over their outer races. This set-up 

Fig 71: Improved curtain-siaer try Montracon 
covered by Patent GB2291095A 

Fig 68: Improved cargo retention system suggested 
by Boalloy Industries, covered in Patent 

Fig 69: Easy-access curtainsider by Cartwright 
Freight Systems (GB 2185715A) 

Fig 70: M & G Tankers & Trailers have suggested 
hinged post sections, in Patent proposal, 



Fig 72: Simplified curtainsider support post in 
Patent GB2209712A by John J Cameron 

Fig 73: Compact sliding door gear by Bedwas 
Bodyworks, in Patent GB2203184A 

Fig 75: Upper-deck levelling achieved by the UK 
Lift Company, covered in Patent GB2299791A 

binding of the rollers when the curtain is displaced 
sideways by the operator during pulling. Each hanger 
bracket is supported by two or more rollers. 

A simplified support post for curtainsider CV 
bodies is proposed in the design of Fig 71. The 
cumbersome toggle linkage provided at the lower 
end of the column is dispensed with. The upper end 
slides on the cant rail, as normal practice, but the 
lower end locks to the side-rave, by means of a simple 
spigot linkage shown here. 

A compact sliding door gear system is offered 
in the design of Fig 73. The difficulty is overcome 
here of accommodating two ISO pallets lengthwise 
across the platform of a goods vehicle — within the 
maximum legal vehicle width and the clearance 
envelope of the door gear. This design uses a ‘master 7 
plug-and-slide door suspended from a carriage and 
swing-arm assembly. 

Tail-lift in a door is the result of the ingenious 
design of Fig 74. A tail-lift platform which doubles as 
a part-closure to prevent cold air fallout from refrig- 
erated vehicles is raised and lowered by carriages in 
the rear doors of the body in this design. The body 
comprises a number of refrigerated compartments 
and a separate folding canopy can temporarily close 
the upper ones. 

Levelling a lifting deck is the object of the Fig 
75 design. Existing mechanisms used to raise the 
decks of goods vehicle bodies, to achieve two-tier 
stowage of the payload, suffer from the moving deck 
becoming unlevel as it is raised on jacks. In this 
design movement of the deck is synchronized by the 
interaction of two toothed racks mounted on the body 
and toothed wheels mounted on the moving deck. 
Connection of the wheels by shaft or endless chain 
ensures the level movement of the deck regardless of 
ram motions. 

Folding catwalk rail is covered by the design of 
Fig 76. Added safety for road tanker operators is 
provided by this pivoting catwalk rail. When access 
to the top of the tanker is not required the rail lays flat 
across the catwalk to prevent ingress by unwanted 
intruders. When legitimate access is required hy- 
draulic rams raise the rail to an upright position while 
interlocks ensure the vehicle’s brakes stay on, in this 

Perhaps the ultimate in cab access is offered by 
the Fig 77 design. Overcoming the ingress and exit 
restrictions for driver and operators of crew-cabbed 



commercial vehicles is the objective, particularly 
suited to vehicles with front overhead loaders. The 
driver’s seat and controls are centrally mounted 
while those for the crew are mounted above and 
behind, on either side of the cab and over the wheel 
housings. Doors of 1.9 metres in height open in- 
wardly to prevent pavement obstruction. The floor is 
inclined upwards at the front and level at the back and 
provides cross cab access ahead of the engine com- 


bus/ambulance design 

Rethinking the ambulance configuration is behind 
the design of Fig 78. Emergency transport of accident 
victims in heavy traffic was the impetus behind an 
interesting vehicle to handle urban emergencies dis- 
cussed at a recent Autotech congress. Authors Da 

controlled by the driver, or ‘rider’ in a semi motor- 
cycle situation, sitting behind the single steering 
front wheel of the vehicle. Retractable outrigger 
wheels would also be used to provide cornering 
stability in less dense traffic conditions. Translucent 
panelling around the patient compartment would 
reduce the claustrophobic effect as would careful 
colour selection for the interior. 

A joint venture between MAN and Voith has 
resulted in the NL 202 DE low floor concept city bus. 
Fig 79, designed to carry 98 passengers at a maxi- 
mum speed of 70 km/h. No steps are involved at any 
of the entrances, which lead directly to a completely 
level deck height of between 317 to 340 mm. The 
rear-mounted horizontally positioned diesel engine 
allows fitment of a bench seat at the rear of the bus; 
it drives a generator with only electrical connection 

Silva and Miranda of Lisbon Technical University 
explained that the key design factor was provision of 
an efficient life condition support for the patient in a 
vehicle with exceptional traffic mobility. That in- 
volved a highly compact layout yet a comfortable 
posture and space for medical assistance to the pa- 
tient in transit. The short journeys inherent in the 
duties of the vehicle make it an attractive proposition 
for electric traction and in the proposed design the 
patient would be positioned above the battery con- 
tainer. Twin rear wheels would be mounted on an 

Fig 76: 
Pivoting cat- 
walk rail 

oscillating axle, the ‘banking’ of which would be 

proposed by 

: safewaik ' 

l Ratings in 

t sr.!— Patent GB 

! 2203390A 

'A D 

hrm v i 

rM;J L_ 


F Bitwnc* 
G Mo««t 

Fig 77; Short-haul ambulance 

H Drr»« Foil 
I • Dnv« Coatioli 

Fig 77: Easy-access cab by Britannia Trucks 



Fig 79a: NL 202 DE 
concept MAN bus 
with propulsion 
system detail below 

to the Voith transverse-flux wheel motors, which 
drive the wheels through two-stage hub-reduction 
gear sets. The diesel is rated at 127 kW and the 
generator at 135 kW; the controller is of the IGBT 
converter type and also developed by Voith. It pro- 
vides a differential action to the wheel motors on 
cornering. Permanent magnet synchronous wheel- 
motors are rated at 57 kW and have a maximum speed 
of 2500 rev/min; see Table 1 in Fig 79a. The bus is 12 
metres long and has water-cooling for its generator, 
converters and wheel motors. As well as providing 
virtually jerk-free acceleration, the drive system is 
seen by MAN as providing the possibility of four- 
wheel drive on future articulated buses to improve 
traction and stability in slippery road conditions. The 
term ‘transverse flux motor’ refers to the means used 
to guide the magnetic flux in the stator; this is new to 
inverter-supplied PM types and involves a novel 
collector configuration. Double-sided magnetic force 
generation is also new and involves a patented double 
air gap construction having high idling inductances 
and force densities up to 120kNm/m 2 , with relatively 
low losses. 

A new control process permits operation of the 
motor in a field-weakening type mode, in spite of 
PM excitation. The generator is almost identical in 
concept but involves no field weakening. Each has 
concentric construction of permanent magnets, ro- 
tor/stator soft-iron elements and stator winding; see 
Fig 79a. Armature elements are U-shaped cut strip- 
wound core sections, embedded in the ring-shaped 
supporting structures of inner and outer stators. Each 
core surrounds the windings and forms a stator pole 
with its cut surfaces facing the rotor. The latter is 
pot-shaped and positioned between poles of the outer 
and inner stators. In the stator pole region it com- 
prises magnet and soft iron element while in the wind- 
ing region a ring of GRP serves as the connecting 

TFM wheel motor TFM generator 


57 kW 

135 kW 

Rated saeed 

735 rev ./min. 

1 ,750 rev ./min. 

Approx, max. speed 

2,500 revymin. 

2,400 rev ./min. 

Max. fundamental frequency of stator 

1 ,350 Hz 


Rated torque 

740 Nm 

740 Nm 

Approx, max. torque 

1,050 Nm 

740 Nm 

Approx, torque conversion 


Power/weight ratio 

1.8 kg/kW 

0.9 kg/kW 



Fig 81: Aero-style luggage rack designed by 
Ikarus covered by Patent GB2209140A 

element. The inverters supply the motors with sinu- 
soidal currents and voltages until the nominal oper- 
ating point is reached; operating frequency is 10 kHz. 
In field-weakening mode the induced voltage exceeds 
intermediate circuit voltage and only square wave 
voltages are supplied to the motor. Power output then 
remains constant and the operating frequency equals 
the fundamental motor frequency. A large speed ra- 
tio, 1.5:1, is thus possible. 

An electric city-bus designed for low drag is 
seen in the Centro concept midibus from Capoco 
Design, Fig 80. It is intended to show the practicality 
of ‘available now’ electric drive technology for city 
centre operation. The 7.3 metre long vehicle would 
operate at up to 40 kph and have a range of 40 miles. 
Energy consumption is reduced by the vehicle's drag 
coefficient of just 0.31 and low rolling-resistance 
tyres of modest section profile. A lightweight struc- 
ture comprises mechanically fastened aluminium 

1 ) EGR electromagnetic control valve 

2) Electronic control unit 

3) Accelerator position reinforcement 

4) Brake pedal switch 

5) Clutch pedal switch 

6) EGR valve 

7) Thermostarter 

8) Thermostarter feed valve 

9) Instrumented injector 

10) Water temperature sensor 

1 1 ) Engine speed sensor 

12) Vehicle speed sensor 

13) Water in fuel sensor 

14) Water flow and temperature sensor 

15) Engine stalling solenoid 

16) Engine advance adjustment 

17) Fuel flow and temperature adjustment 

18) Injection pump. 

Fig 79 b: Performance characteristic and 
power-unit environmental control system 



alloy extrusions covered by a single-piece roof mould- 
ing. In full standee form up to 56 passengers can be 
carried and deck height is just 290 mm. Shown also 
in Fig 80 is the economical packaging of batteries 
and drive motor, with near vertical drive axis. Width 
of the vehicle is 2380 mm and height 2550 m, on a 
4050 mm wheelbase. 

An aero-style bus saloon assembly follows 
from modular elements used in this patented roof 
liner and luggage rack assembly, Fig 81, for touring 
coaches. The arrangement involves zed-shaped brack- 
ets screwed to the roof frame which secure headlin- 
ing, ventilating ducts and luggage shelves. The shelf 
elements incorporate a profiled channel, for housing 
light fittings, which is closed by a snap-fit translucent 

Fig 80: Capoco Centro electric bus 

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Chapter 1 

1 Sears, K., Automotive engineering: strategic over- 
view, Steel Times, Vol. 2. No.l, 1997 

2 Matthews and Davies, Precoated steels develop- 
ment for the automotive industry, Proc. I.Mech.E., 
Vol. 211., Part D, 1997 

3 Rink and Pugh, The perfect couple — metal/plastic 
hybrids making effective use of composites, IBCAM 
Conference, 1997 

4 Merrifield, R., Instrument panel structural concepts 
which integrate functions utilizing injection moulded 
plastic components. Paper C524/120/97, Autotech 
Congress, 1997 

5 Ashby, M., Material Property Charts, Perform- 
ances Indices, Section 4, Material Selection and 
Design, 20th edition , ASM International, 1997 

Chapter 2 

1 Bass et al, A system for simulating structural 
intrusion, Proc. I.Mech.E., Vol. 211., Part D, 1997 

2 Yim and Lee, Design optimization of the pillar joint 
structures using equivalent beam modelling tech- 
nique, SAE Paper 971544, part of P-308 

3 Hardy, M., Automotive modelling and NVH semi- 
nar, Proc. I.Mech.E., 1997 

4 Hargreaves, J., I .Mech.E .Vehicle noise and vibra- 
tion conference, 1998 

5 Lawrence and Hardy, Development and use of 
pedestrian impactors to reduce the injury potential of 
cars, paper C5 24/049/97, Autotech 1997 

6 Coleman and Harrow, Car design for all, IMechE 
seminar , 1997 

7 Delphi Automotive, Adaptive restraint technolo- 
gies, Vehicle Engineer, December 1998 

Chapter 3 

1 Evans and Blaszczyk, A comparative study of the 
performance and exhaust emissions of a spark igni- 
tion engine fuelled by natural gas and gasoline, 
Proc. I.Mech.E., Vol. 211, Part D, 1997 

2 Perkins and Penny, Design options and perform- 
ance characteristics for 0.25-0.3 litre/cylinder HSDI 
diesel engines, Euro 4challenge: future technology 
and systems, IMechE Seminar, 1997 

3 Sadler et al., Optimization of the combustion 
system for a direct-injection gasoline engine using a 

high speed in-cylinder sampling valve, Euro 4chal- 
lenge: future technology and systems, I.Mech.E. 
Seminar, 1997 

4 Zhou and Qian, Development of a modified diesel 
engine cycle, Proc. I.Mech.E., Vol 212, Part D, 1998 

5 Cains et al.. High dilution combustion through 
axial and barrel swirl, I.Mech.E. Seminar Publica- 
tion: Automotive Engines and Powertrains, 1997 

6 Bassett et al., A simple two-state late intake valve 
closing mechanism, Proc IMechE, Vol 211, Part D, 

7 Woods and Brown, Row area of multiple poppet 
valves, Proc IMechE, Vol 210, Part D, 1996 

8 Ohashi et al., Honda’s 4 speed all clutch to clutch 
automatic transmission, SAE Paper 980819 

9 Abo et al., Development of a metal belt-drive CVT 
incorporating a torque converter for use with 2-litre 
class engines, SAE Paper 980823 

10 Ahluwalia et al, The new high torque NVT-750 
manual transaxle, SAE Paper 980828 

1 1 Turner and Kelly, A transmission for all seasons. 
Advanced vehicle transmission and powertrain con- 
ference, I.Mech.E., 1997 

12 1998 SAE software based on book: Gillespie, G, 
Fundamentals of vehicle design, SAE, 1992 

13 Gadola and Cambiaghi, MMGB: a computer- 
based approach to racing car suspension design, A TA, 
Vol. 47., no. 6/7., 1994 

14 Ellis, J, Vehicle handling dynamics, Mechanical 
Engineering Publications, 1994 

15 Potter et al, Assessing ‘road-friendliness’: a 
review, Proc. I.Mech.E., Vol. 211., Part D, 1997 

Chapter 4 

la. Bodoni-Bird, C, What can seamless electro- 
mechanical vehicles learn from Nature?, 96C001, 

lb. Miyata et al, Engine control by ion density 
analysis, 96C003 

lc. Pinkos and Shtarkman, Electronically controlled 
smart materials in active suspension systems, 96C004 

l d. Hatanaka and Noro, New approach for intelligent 
steering system development, 96C007 

le. Milbum, S, Integration of advanced functions 
into electric drivetrains, 96C050 

...from the bound volume of proceedings of the 1996 
Convergence Transportation Electronics Association 

2a Kuragaki et al., An adaptive cruise control using 
wheel torque management technique, 980606 



2b Olbrich et al, Light radar sensor and control unit 
for adaptive cruise control, 980607 
...from the bound volume of proceedings of the ITS 
Advanced controls and vehicle navigation systems 
seminar at the 1998 SAE Congress 

3 Bolton, W., Essential mathematics for engineer- 
ing, Heinemann 

4 Jones et al, HYZEM — a joint approach towards 
understanding hybrid vehicle introduction into 
Europe, Proceedings of the IMechE Combustion 
engines and hybrid vehicles conference, 1998 

5 Friedmann et al. , Development and application of 
map-controlled drive management for a BMW parallel 
hybrid vehicle, SAE Special Publication SP1331, 

6 Nagasaka et al.. Development of the hybrid/battery 
ECU for the Toyota Hybrid System, SAE Special 
Publication SP1331, 1998 

7 Saito et al., Super capacitor for energy recycling 
hybrid vehicle, Convergence 96 proceedings 

8 Qriguchi et al.. Development of a lithium-ion 
battery system for EVs, SAE paper 970238 

Chapter 5 

1 Bowsher , G, Braking and traction at supersonic 
speeds, Special Vehicle Engineer, January 1998 

Chapter 7 

1 SAE papers 870147/8 

2 I.Mech.E .Autotech paper C427/40/108 

3 Proc. I.MechE., Vol 206, Aerodynamics of Grand 
Prix cars 

4 I.Mech.E .Autotech Paper C427/6/032 





Access to door hardware 156 
Acoustic model of vehicle cabin 21 
Adaptive restraint technologies 31 
Advanced car and truck engines 120 
Advanced suspension linkage 133 
Aero-style luggage rack 170 
Aerodynamics 161 
Aged 50+ coupe 27 
Airbag module as horn-push 152 
Al 2 concept car 3 

Alternative injector installations 41 
Alternative valvetrain layouts 42 
Anti-dive motor cycle suspension 137 
Ashland Pliogrip fast-cure adhesive 8 
ASI passenger-side low-mount airbag 31 
Auto-adjust towbar 142 
Auto-steer for trailer 131 
Automatic cruise control, ACC 78 
Automatic drive selection 129 
Automation of handling tests 90 
Automotive electronics maturity 70 


Bayer Durethan BKV polymer composite 4 

BMW for its rear suspension, 850i coupe 133 

BMW parallel hybrid drive 85 

Body shell integrity 145 

Body structure and systems. Ford Focus 105 

Body systems of Freelander 109 

Body-in-white of M-B A-class 98 

Bonnet latch 159 

Boot-stowage of roof 158 

Bosch Motronic MED 7 management system 81 

Brake apportioning for solo or coupled mode 141 

Braking system of M-B A-class 101 

Braking systems 139 

Bus and ambulance design 168 


Cabin acoustic model 22 

Calliper and disk assembly, Thrust SSC 1 14 

Cambridge Consultants composite structures 148 

Capoco Centro electric bus 170 

Car body systems 151 

Cargo retention system 166 

Chassis systems 133 

Chassis/body shell elements 147 

Child safety seat 151 

Chrysler China car 7 

Clock enable circuit 80 

CNG's advantage over gasoline 39 

Collins CMC scotch-yoke engine 127 

Compact door latch 158 

Compact sliding door gear 167 

Compliance representation 64 

Constant-pressure cycle 121 

Constant pressure cycle piston modification 122 

Constant-pressure cycle: the future for diesels? 45 

Control Blade rear suspension 104 

Control strategies for CVT 56 

Control strategy for 4- wheel steer 138 

Controlled -collapse trim 151, 160 

Controlled collapse 144 

Controlling restraint deployment 33 

Conventional and electric drivetrain 75 

Cos worth Engineering’s MBA engine 118 

Crash severity sensing 31 

Cross-bolting of main-bearing caps 1 1 8 

C V chassis-cab configuration 1 63 

CV systems 163 

CV-joint packaging 131 

CVT engine torque curve matching 56 

CVT for 2-litre engined vehicles 53 


DAF concept truck 163 

Dallara MMGB suspension design software 63 

Delphi E-STEER electronic steering 82 

Dennis Rapier space-frame chassis 164 

Design for the disabled 27 

Diesel engine with oil-cushioned piston 45 

Digital circuits for computation 79 

Direct injection gasoline 43 

Door seals 155 

Doors, windows and panels 155 
Double-level floor of Mercedes A-class 97 
Drive and steer systems 128 
Drive strategy for influencing factors 57 
Driverless taxi 29 
Drivetrain control 75 


Easy-access cab 168 
Easy-access curtainsider 166 



Easy-assemble brake booster 142 
Easy-change CV-UJ assembly 130 
Electric double-layer capacitors (EDLCs) 88 
Electro-rheological magnetic (ERM) fluids 72 
Electromagnetic braking 141 
Electromagnetic valve actuation 122 
Electronic brake actuation system 139 
Electronic control of electric steering 73 
Electronic Stability Programme on Focus 103 
Emissions of small HSDI diesels 40 
Energy storage initiatives 88 
Engine developments 118 
Engine force balancer for motorcycle 126 
Engine refinement 124 
Engine temperature management 121 
Engine-coolant airflow 132 
Equilibrium lateral acceleration 59 
European 13 Mode test 39 
Exhaust fixing 161 


Fail-safe brake actuator 141 

Fatigue life of valve gear 125 

Fatigue performance of Fastriv system 14 

Fibre-optic light division system 161 

Filtering diesel particulates 127 

Fluid-actuated anti-roll control 136 

Flywheel battery 128 

Flywheel to crankshaft connection 126 

FMVSS impact energy absorption levels 7 

Footwell deformation in vehicle impact 16 

Ford Focus 102 

Freer moving bogie suspension 136 

Front and rear wheels, Thrust SSC 112 

Front brake, Thrust SSC 115 

Fuel delivery system 161 

Fuel injection rethought 122 

Fuel stratification 44 

Fuji Industies ELCAPA hybrid vehicle 89 


Gas storage requirement 39 
Gear train schematics 52 


Handling software example 60 
Head rest adjuster 154 
Headway sensor design 77 

Height-adjust for trailer coupling 142 
High activity, homogeneous charge concept 47 
High-torque manual transmission 54 
Honda automatic transmission 53 
Honda intelligent steering system 74 
Hybrid drive prospects 84 
Hydraulically sprung connecting rod 45 


Ideal cycle efficiency 46 

Intake valve disablement 47 

Integrated air/fuel induction system 122 

Integrated engine/transmission 43 

Integrated latch/lock system by Bosch 157 

Integration of clutch servo actuator parts 142 

Interior noise analysis 20 

Interior sound pressure distribution 23 

Intrusion simulation mechanism 17 

Ion current response 70 


Kyoto tram 30 


Land Rover Freel ander 108 

Late intake valve closing 48 

Lateral acceleration of tractor/trailer 61 

LD-SR1M composite panel construction 162 

Leg injuries 33 

Light-alloys development 3 

Liquified natural gas (LNG) 40 

LNG-powered ERF truck 39 

Logic circuits 79 

Lotus Elan structure 149 

Lotus Engineering Jewel Project engine 120 

Low-cost supercharging 122 


Magnesium Association Design Award 3 

MAN diesel pollution control 1 26 

MAN and Voith NL 202 DE low floor city bus 168 

Material property charts/performance indices 10 

Mazda punt structure 150 

Mechanical systems on Free] ander 110 

Mechanics of roll-over 58 

Mercedes A-class 96 

Metal/plastic systems 4 

Millbrook VTEC facility 38 

Mobility for all 27 



Multi input/output gearbox 129 
Multi-node suspension model 62 
Multiple valve arrangements 49 


Navigation system advances 76 
Necked hub member 138 
New Venture Gear NVT-750 gearbox 55 
Nissan CVT configuration 54 
Nissan CVT for 2-litre cars 53 
Non-stick front forks 136 


Occupant protection in side impact 146 

Occupant restraint 151 

Occupant sensing 31 

Optimized re-charge strategy 86 

Optimizing unsprung mass with wheel travel 133 


Parallel hybrid drive mechanism 85 

Part-load control by late intake valve closing 48 

Particle orientation in rheological fluid 72 

Particle size analysis 37 

Particulate size distribution 37 

Passive Anti-Theft System (PATS) 107 

Pedestrian protection device 160 

Pedestrian protection in impact 24 

PEM fuel cell, natural gas fuelled 89 

Personal Productivity info-technologies 81 

Pillar to rail joint stiffness 18 

Plastic structural beam 6 

Plenum and port throttles 1 19 

Plunge joint 130 

Poppet valve effective area comparison 50 
Poppet valve half-angles/radius-ratios 5 1 
Porshe suspension sub-frame 148 
Power interaction diagram for hybrid drive 86 
Powertrain on Ford Focus 104 
Powertrains: the next stage? 36 
Primary safety system of M-B A-class 99 
Production hybrid-drive control system 87 
Proprietary control system advances 81 
Pyrotechnic actuated venting 33 


Quicker catalyst light-up 127 


RACD damper unit 73 

Racing clutch 129 

Rapid-actuation diff-lock 129 

Rear brake. Thrust SSC 116 

Recycled PET, and prime PBT, for sun-roof parts 7 

Reduced-emission systems 124 

Refinement of individual systems, Ford Focus 102 

Return-load tanker 165 

Road guidance for drowsy drivers 77 

Road traffic NOx emissions 38 

Road-friendliness, a current review 66 

Roll effects 65 

Roll reactions on suspended vehicle 58 

Roll response to step input 59 

Rover Asymmetric Combustion Enhancement 47 


Sachs variable valve timing 125 

Sarich 2-stroke 124 

Scania turbo-compound diesel 120 

Seamless electro-mechanical vehicle 70 

Seat adjustment mechanism 153 

Seat of Mercedes 300 SL 153 

Seat-belt pretensioners 1 5 1 

Secondary safety systems 96 

Self-pierce riveting 13 

Self-steering in forward and reverse 138 

Set-speed down-hill braking 142 

Shopping ferry 30 

Short-haul taxi 28 

Side airbag within seat back 152 

Side impact protection structure 145 

Sidemember construction novelty 147 

Sill and quarter panel joint 146 

Simplified level sensing 1 60 

Sliding door drive 159 

Smart materials for suspension control 7 1 

Soft-mounted cab tilt 164 

Specific stiffness property chart 1 1 

Specific strength property chart 12 

Speeding engine warm-up 127 

Split tailgate 155 

Steel durability and structural efficiency 2 
Steer fight in braking 135 
Steering assist mechanism 82 
Steering column immobilizer 137 
Steering robot 91 



Steering system on Freelander 111 
Steering-castor adjustment 137 
Storing swirl energy 122 
Strengthening a convertible body 148 
Structural-coloured fibre has body trim potential 8 
Structural design in polymer composites 6 
Structural systems 144 
Structure of Freelander 108 
Supercapacitors for hybrid drive 88 
Super-element in FEA 20 
Sure-fire gear engagement 129 
Suspension and steering linkage analysis 62 
Suspension compliance conflict 135 
Suspension compliance work of John Ellis 64 
Suspension development 133 
Suspension geometry control 135 
Suspension supports of M-B A-class 100 
Swivel pin bearing pre-stressing 137 
Synthetic urban drive cycle 84 
Systems integration effect on sensors 70 


Tail-lift in door 167 

Taxis and people movers 28 

TEC mobility system 81 

Textron Fastriv system 13 

Through-bolt and monoblock configurations 43 

Thrust SSC land-speed jet car 112 

Tiller steering 137 

Torque reaction system for steering robot 92 

Torque roll axis engine mounts on Focus 107 

Toyota Hybrid System 87 

Transient roll-over 59 

Transmission design trends 52 

Transverse spring composite suspension 133 

Tribus guided vehicle system 128 

Trim and fittings 159 

TRL knee joint model 24 

TRL leg-form impactor 25 

Truck and bus emissions 38 

Tuned suspension mounts on Focus 107 

Turbocharger waste-gate control 121 

Two-dimensional suspension analysis 64 

Two-stage energy absorption 144 


ULSAB body shell 2 
Urban rickshaw 27 


Valve arrangements for engine efficiency 47 

Variable cube body 165 

Variable geometry engine intake 125 

Variable valve timing 125 

Vehicle body FEM 23 

Visualization of in-cylinder flow 44 

VVT for Rover K-series 124 


Wheeled battle-tank 132 
Winder mechanism 156 


Zero-offset geometry 103